Centrifugal heat transfer engine and heat transfer systems embodying the same

ABSTRACT

A heat transfer engine having cooling and heating modes of reversible operation, in which heat can be effectively transferred within diverse user environments for cooling, heating and dehumidification applications. The heat transfer engine of the present invention includes a rotor structure which is rotatably supported within a stator structure. The stator has primary and secondary heat exchanging chambers in thermal isolation from each other. The rotor has primary and secondary heat transferring portions within which a closed fluid flow circuit is embodied. The closed fluid flow circuit within the rotor has a spiraled fluid-return passageway extending along its rotary shaft, and is charged with a refrigerant which is automatically circulated between the primary and secondary heat transferring portions of the rotor when the rotor is rotated within an optimized angular velocity range under the control of a temperature-responsive system controller. During the cooling mode of operation, the primary heat transfer portion of the rotor carries out an evaporation function within the primary heat exchanging chamber of the stator structure, while the secondary heat transfer portion of the rotor carries out a condenser function within the secondary heat exchanging chamber of the stator. During the cooling mode of operation, a vapor-compression refrigeration process is realized by the primary heat transfer portion of the rotor performing an evaporation function within the primary heat exchanging chamber of the stator structure, while the secondary heat transfer portion of the rotor performs a condenser function within the secondary heat exchanging chamber of the stator. During the heating mode of operation, a vapor-compression refrigeration process is realized by the primary heat transfer portion of the rotor performing a condenser function within the primary heat exchanging chamber of the stator structure, while the secondary heat transfer portion of the rotor performs an evaporation function within the secondary heat exchanging chamber of the stator. By virtue of the present invention, a technically feasible heat transfer engine is provided which avoids the need for conventional external compressors, while allowing the use of environmentally safe refrigerants. Various embodiments of the heat transfer engine are disclosed, in addition to methods of manufacture and fields and applications of use.

RELATED CASES

[0001] This is a Continuation of copending Application Ser. No.09/317,055 filed May 24, 1999, which is a Continuation of ApplicationSer. No. 08/725,648 filed Oct. 1, 1996, now U.S. Letters Patent5,906,108, which is a Continuation of copending Application Ser. No.08/656,595 filed May 31, 1996, which is a Continuation of ApplicationSer. No. 08/391,318 filed Feb. 21, 1995, which is a Continuation ofApplication Ser. No. 08/175,485 filed Dec. 30, 1993, which is aContinuation of Application Ser. No. 07/893,927 filed Jun. 12, 1992,each of said Applications being assigned to and commonly owned byKidwell Environmental, Ltd., Inc. of Tulsa, Oklahoma and incorporatedherein by reference in its entirety.

BACKGROUND OF INVENTION

[0002] 1. Field of the Invention

[0003] The present invention relates to a method of and apparatus fortransferring heat within diverse user environments, using centrifugalforces to realize the evaporator and condenser functions required in avapor-compression type heat transfer cycle.

[0004] 2. Brief Description of the State of the Prior Art

[0005] For more than a century, man has used various techniques fortransferring heat between spaced apart locations for both heating andcooling purposes. One major heat transfer technique is based on thereversible adiabatic heat transfer cycle. In essence, this cycle isbased on the well known principle, in which energy, in the form of heat,can be carried from one location at a first temperature, to anotherlocation at a second temperature. This process can be achieved by usingthe heat energy to change the state of matter of a carrier fluid, suchas a refrigerant, from one state to another state in order to absorb theheat energy at the first location, and to release the absorbed heatenergy at the second location by transforming the state of the carrierfluid back to its original state. By using the reversible heat transfercycle, it is possible to construct various types of machines for bothheating and/or cooling functions.

[0006] Most conventional air conditioning systems in commercialoperation use the reversible heat transfer cycle, described above. Ingeneral, air conditioning systems transfer heat from one environment(i.e. an indoor room) to another environment (i.e. the outdoors) bycyclically transforming the state of a refrigerant (i.e. working fluid)while it is being circulated throughout the system. Typically, the statetransformation of the refrigerant is carried out in accordance with avapor-compression refrigeration cycle, which is an instance of the moregenerally known “reversible adiabatic heat transfer cycle”.

[0007] According to the vapor-compression refrigeration cycle, therefrigerant in its saturated vapor state enters a compressor andundergoes a reversible adiabatic compression. The refrigerant thenenters a condenser, wherein heat is liberated to its environment causingthe refrigerant to transform into its saturated liquid state while beingmaintained at a substantially constant pressure. Leaving the condenserin its saturated liquid state, the refrigerant passes through athrottling (i.e. metering) device, wherein the refrigerant undergoesadiabatic throttling. Thereafter, the refrigerant enters the evaporatorand absorbs heat from its environment, causing the refrigerant totransform into its vapor state while being maintained at a substantiallyconstant pressure. Consequently, as a liquid or gas, such as air, ispassed over the evaporator during the evaporation process, the air iscooled. In practice, the vapor-compression refrigeration cycle deviatesfrom the ideal cycle described above due primarily to the pressure dropsassociated with refrigeration flow and heat transfer to or from theambient surroundings.

[0008] A number of working fluids (i.e. refrigerants) can be used withthe vapor-compression refrigeration cycle described above. Ammonia andsulfur dioxide were important refrigerants in the early days ofvapor-compression refrigeration. In the contemporary period, azeotropicrefrigerants, such as R-500 and R-502, are more commonly used.Halocarbon refrigerants originate from hydrocarbons and include ethane,propane, butane, methane, and others. While it is a common practice toblend together three or more halogenated hydrocarbon refrigerants suchas R-22, R-125, and R-290, near-azeotropic blend refrigerants sufferfrom temperature drift. Also, near azeotropic blend refrigerants areprone to fractionation, or chemical separation. Hydrocarbon based fluidscontaining hydrogen and carbon are generally flammable and therefore arepoorly suited for use as refrigerants. While halogenated hydrocarbonsare nonflammable, they do contain chlorine, fluorine, and bromine, andthus are hazardous to human health.

[0009] Presently, the main refrigerants in use are the halogenatedhydrocarbons, e.g. dichlorodifluoromethane (CCL2F2), commonly known asR-12 refrigerant. Generally, there are three groups of usefulhydrocarbon refrigerants: chlorofluorocarbons, (CFCs),hydrochlorofluorocarbons, (HCFCs), which are created by substitutingsome or all of the hydrogen with halogen in the base molecule.Hydrofluorocarbons, (HFCs), contain hydrogen, fluorine, and carbon.However, as a result of the Montreal Protocol, CFCs and HCFCs are beingphased out over the coming decades in order to limit the production andrelease of CFC's and other ozone depleting chemicals. The damage toozone molecules (O₃) comprising the Earth's radiation-filtering ozonelayer occurs when a chlorine atom attaches itself to the O₃ molecule.Two oxygen atoms break away leaving two molecules. One molecule isoxygen (O₂) and the other is chlorine monoxide molecule (CO). Thechlorine monoxide is believed by scientists to displace the ozonenormally occupying that space, and thus effectively depleting the ozonelayer.

[0010] While great effort is being expended in developing newrefrigerants for use with machines using the vapor-compressionrefrigeration cycle, such refrigerants are often unsuitable forconventional vapor-compression refrigeration units because of theirincompatibility with existing lubricating additives, and the levels oftoxicity which they often present. Consequently, existingvapor-compression refrigeration units are burdened with a number ofdisadvantages. Firstly, they require the use of a mechanical compressorwhich has a number of moving parts that can break down. Secondly, theworking fluid must also contain oil to internally lubricate thecompressor. Mineral oil has been used in refrigeration systems for manyyears, and alternative refrigerants like hydrofluorocarbons (HFC)require synthetic lubricants such as alkylbenzene and polyester. Thisuse of such lubricants diminishes system efficiency. Thirdly, existingvapor-compression systems require seals to prevent the escape of harmfulrefrigerant vapors. These seals can harden and leak with time. Lastly,new requirements for refrigerant recovery increase the cost of avapor-compression unit.

[0011] In 1976, Applicant disclosed a radically new type ofrefrigeration system in U.S. Pat. No. 3,948,061, now expired. Thisalternative refrigeration system design eliminated the use of acompressor in the conventional sense, and thus many of the problemsassociated therewith. As disclosed, this prior art system comprises arotatable structure having a hollow shaft with a straight passagetherethrough, and about which a closed fluid circuit is supported. Theclosed fluid circuit is realized as an assemblage of two spiral tubularassemblies, each consisting of first land second spiraled tube sections.The first and second spiraled tube sections have a different number ofturns. A capillary tube, placed between the condenser and evaporatorsections, functions as a throttling or metering device. When therotatable structure is rotated in a clockwise direction, one end of thetube assembly functions as a condenser, while the other end thereoffunctions as an evaporator. As disclosed, means are provided fordirecting separate streams of gas or liquid across the condenser andevaporator assemblies for effecting heat transfer operations with theambient environment.

[0012] In principal, the refrigeration unit design disclosed in U.S.Pat. No. 3,948,061 provides numerous advantages over existingvapor-compression refrigeration units. However, hitherto successfulrealization of this design has been hindered by a number of problems. Inparticular, the use of the capillary tube and the hollow shaft passagecreate imbalances in the flow of refrigerant through the closed fluidflow circuit. When the rotor structure is rotated at particular speeds,there is a tendency for the refrigerant fluid to cease flowingtherethrough, causing a disturbance in the refrigeration process. Also,when using this prior art centrifugal refrigeration design, it has beendifficult to replicate the refrigeration effect with reliability, andthus commercial practice of this alternative refrigeration system andprocess has hitherto been unrealizable.

[0013] Thus, there exists a great need in the art for an improvedcentrifugal heat transfer engine, which avoids the shortcomings anddrawbacks thereof, and allows for the widespread application of such analternative heat transfer technology in diverse applications.

OBJECTS OF THE PRESENT INVENTION

[0014] Accordingly, it is a primary object of the present invention toprovide an improved method of and apparatus for transferring heat withindiverse user environments using centrifugal forces to realize theevaporator and condenser functions required in a vapor-compression typeheat transfer cycle, while avoiding the shortcomings and drawbacks ofprior art apparatus and methodologies.

[0015] A further object of the present invention is to provide suchapparatus in the form of a centrifugal heat transfer engine which, byeliminating the use of mechanical compressors, reduces the introductionof heat into the system by the internal moving parts of conventionalmotor driven compressors, and energy losses caused by refrigerationlubricants used to lubricate the moving parts thereof.

[0016] A further object of the present invention is to provide acentrifugal heat transfer engine that contains the refrigerant within aclosed system in order to avoid leakage, yet being operable with a widerange of refrigerants. having A further object of the present inventionis to provide a centrifugal heat transfer engine having a rotorstructure with a closed, fluid circulating system that contributes to adynamic balance of refrigerant flow.

[0017] A further object of the present invention is to provide acentrifugal heat transfer engine having a rotor structure embodying afluid circulation system which, when rotated direction in a firstdirection, has a first portion that functions as a condenser and asecond portion that functions as an evaporator to provide arefrigeration unit, and when the direction of the rotor structure isreversed, the first portion functions as an evaporator and the secondportion functions as a condenser to provide a heating unit.

[0018] A further object of the present invention is to provide acentrifugal heat transfer engine that either condenses or evaporates achemical refrigerant as it is passed through a plurality of helicalpassageways which are part of its rotor structure.

[0019] A further object of the present invention is to provide acentrifugal heat transfer engine which provides a simple apparatus forcarrying out a refrigeration cycle without the necessity for compressorsor other internal moving parts that introduce unnecessary heat into therefrigerant.

[0020] A further object of the present invention is to provide acentrifugal heat transfer engine which does not require refrigerantcontamination with an internal lubricant, and thus permits therefrigerant to function at optimum heat transferring quality.

[0021] A further object of the present invention is to provide acentrifugal heat transfer engine having a temperature responsivetorque-controlling system in order to maintain the angular velocity ofthe rotor structure within prespecified operating range, and thusmaintain the flow of refrigerant through the fluid circulating system ofthe rotor structure.

[0022] A further object of the present invention is to provide such acentrifugal heat transfer engine with a rotatable structure containingthe self-circulating fluid circuit having a bidirectional throttlingdevice placed between the condenser section and the evaporator sectionof the fluid circuit.

[0023] A further object of the present invention is to provide such abidirectional throttling device for controlling the flow rate of liquidrefrigerant into the evaporation length of the evaporator section of therotor structure, and the amount of pressure drop between the liquidpressurization length and the evaporation length during a range of axialvelocities (RPM) of the rotor structure.

[0024] A further object of the present invention is to provide such acentrifugal heat transfer engine, in which the optimum axial velocity isarrived at and controlled by a torque controlling system responsive totemperature changes detected in the ambient air or liquid being treatedusing an array of temperature sensors.

[0025] A further object of the present invention is to provide such acentrifugal heat transfer engine with a spiral passage along the shaftof the rotor structure in order to cause vapor-compression as it drawsthe heavy refrigerant vapor from the evaporator to the condenser in bothclockwise and counterclockwise directions of rotation.

[0026] A further object of the present invention is to provide such acentrifugal heat transfer engine with a rotor structure having heattransfer fins in order to enhance heat transfer between the circulatingrefrigerant and the ambient environment during the operation of theengine.

[0027] A further object of the present invention is to provide such acentrifugal heat transfer engine, in which the closed refrigerant flowcircuit within the rotor structure is realized as spiraled tubingassembly having spiraled tubular condenser section and a tubularevaporator Section which are both held in position by structuralsupports anchored to the shaft and connected to spiraled tubes.

[0028] A further object of the present invention is to provide such acentrifugal heat transfer engine, in which the rotor structure isconstructed as a solid assembly and the closed refrigerant flow circuit,including its spiral return passageway along the axis of rotation, isformed therein.

[0029] Another object of the present invention is to provide a novelheat transfer engine which can be used to transfer heat within abuilding, home, automobile, tractor-trailer, aircraft, freight train,maritime vessel, or the like, in order to maintain one or moretemperature control functions.

[0030] These and other objects of the present invention will becomeapparent hereinafter and in the Claims to Invention.

SUMMARY OF THE INVENTION

[0031] In general, the present invention provides a novel method andapparatus for transferring heat within diverse user environments, usingcentrifugal forces to realize the evaporator and condenser functionsrequired in a vapor-compression type heat transfer cycle.

[0032] According to a first aspect of the present invention, theapparatus of the present invention is provided in the form of areversible heat transfer engine. The heat transfer engine comprises astator, port connectors, a heat exchanging rotor, torque generator,temperature selector, a plurality of temperature sensors, a fluid flowrate controller, and a system controller.

[0033] The stator housing has primary and secondary heat transferchambers, and a thermal isolation barrier disposed therebetween. Theprimary and secondary heat transfer chambers each have inlet and outletports and a continuous passageway therebetween. A first port connectoris provided for interconnecting a primary heat exchanging circuit to theheat ports of the primary heat transfer chamber, so as to permit aprimary heat exchanging medium to flow through the primary heatexchanging circuit and the primary heat exchanging chamber during theoperation of the heat transfer engine. A second port connector isprovided for interconnecting a secondary heat exchanging circuit to theinlet and outlet ports of said secondary heat transfer chamber, so as topermit a secondary heat exchanging medium to flow through the secondaryheat exchanging circuit and the secondary heat transfer chamber duringthe operation of the reversible heat transfer engine, while the primaryand secondary heat exchanging circuits are in substantial thermalisolation of each other.

[0034] The heat exchanging rotor is rotatably supported within thestator housing about an axis of rotation and having a substantiallysymmetrical moment of inertia about the axis of rotation. The heatexchanging rotor has a primary heat exchanging end portion disposedwithin the primary heat transfer chamber, a secondary heat exchangingend portion disposed within the secondary heat transfer chamber, and anintermediate portion disposed between the primary and secondary heatexchanging end portions. The heat exchanging rotor contains a closedfluid circuit symmetrically arranged about the axis of rotation and hasa return portion extending along the direction of the axis of rotation.

[0035] The primary heat exchanging end portion of the rotor is disposedin thermal communication with the primary heat exchanging circuit, andthe secondary heat exchanging end portion of the rotor is disposed inthermal communication with the secondary heat exchanging circuit. Theintermediate portion of the rotor is physically adjacent to the thermalisolation barrier so as to present a substantially high thermalresistance to heat transfer between the primary and secondary heatexchanging chambers during operation of the heat transfer engine.

[0036] A predetermined amount of a heat carrying medium is containedwithin the closed fluid circuit of the heat exchanging rotor. The heatcarrying medium is characterized by a predetermined heat of evaporationat which the heat carrying medium transforms from liquid phase to vaporphase, and a predetermined heat of condensation at which the heatcarrying medium transforms from vapor phase to liquid phase. Thedirection of phase change of the heat carrying liquid is reversible.

[0037] The function of the torque generator is to impart torque to theheat exchanging rotor and cause the heat exchanging rotor to rotateabout the axis of rotation. The function of the temperature selector isto select a temperature to be maintained along the primary heatexchanging circuit. The function of the temperature sensor is to measurethe temperature of the primary heat exchanging medium flowing throughthe inlet and outlet ports of the primary heat exchanging chamber, andfor measuring the temperature of the secondary heat exchanging mediumflowing through the inlet and outlet ports of the primary heatexchanging chamber. The function of the fluid flow rate controller is tocontrol the flow rate of the primary heat exchanging medium flowingthrough the primary heat exchanging chamber and the flow rate of thesecondary heat exchanging medium flowing through the secondary heatexchanging chamber, in response to the sensed temperature of the heatexchanging medium at either the inlet or outlet port in either theprimary or secondary heat exchanging chambers and to satisfy thetemperature selector setting.

[0038] The function of the torque controller is to control the torquegenerating means in response to the sensed temperature of the heatexchanging medium at either the inlet or outlet port in either theprimary or secondary heat exchanging chambers and the selected operatingtemperature setting.

BRIEF DESCRIPTION OF THE DRAWINGS

[0039] For a more complete understanding of the Objects of the PresentInvention, the following Detailed Description of the IllustrativeEmbodiments should be read in conjunction with the accompanyingDrawings, wherein:

[0040]FIG. 1 is a schematic representation of the first illustrativeembodiment of the heat transfer engine of the present invention, showingthe fluid-carrying rotor structure thereof being rotated about its shaftby a torque generator controlled by a system controller responsive tothe temperatures measured from a plurality of locations about thesystem;

[0041]FIG. 2A is an elevated side view of the fluid-carrying rotorstructure of the first illustrative embodiment of FIG. 1, shown removedfrom the stator portion thereof, and with indications depicting whichfluid carrying tube sections carry out the condenser and evaporatorfunctions respectively, when the rotor structure is rotated in thedirection shown;

[0042]FIG. 2B is a top view of the fluid-carrying rotor structure of thefirst illustrative embodiment of the FIG. 1, shown removed from thestator portion thereof, with indications depicting the location of thethrottling device and rotor shaft coil penetrations;

[0043]FIG. 3 is an elevated side view of the fluid-carrying rotorstructure of the first illustrative embodiment of FIG. 1, shown removedfrom the stator portion thereof, with indications depicting which fluidcarrying tube sections carry out the condenser and evaporator functions,respectively, when the rotor structure is rotated in the directionshown;

[0044]FIG. 4A is an elevated side view of the rotatable support shaft ofthe rotor structure of the first illustrative embodiment of FIGS. 1 and2, showing the spiraled passageway extending therealong and shaft endbearing surfaces machined in the shaft core material;

[0045]FIG. 4B is an elevated cross-sectional side view of the rotatablesupport shaft of FIG. 4A, shown inserted into its shaft cover sleeve andwelded thereto with a bead of weld formed around the circumferencethereof,

[0046]FIG. 5 is an elevated cross-sectional longitudinal view of therotatable support shaft of the rotor structure of the first illustrativeembodiment of FIG. 1;

[0047]FIGS. 6A and 6B are cross-sectional views of the rotatable supportshaft of the rotor structure of the first illustrative embodiment takenalong lines 6A-6A and 6B-6B, respectively, of FIG. 5, showing the mannerin which the end portions of the spiral coil structure are connected tothe spiraled passage formed along the rotatable support shaft of therotor structure of the first illustrative present invention;

[0048]FIG. 7A is a first elevated side view of a support element used tosupport a section of the fluid-carrying spiraled tube portion of therotor structure of the first illustrative embodiment of the presentinvention;

[0049]FIG. 7B is a second elevated side view of the support elementshown in FIG. 7A;

[0050]FIG. 7C is an elevated axial view of one spiral turn of thefluid-carrying spiraled tube portion of the rotor structure of the firstillustrative embodiment of the present invention shown in FIG. 1;

[0051]FIG. 8A is a schematic representation of the heat transfer engineof the first illustrative embodiment of the present invention installedwithin a heat transfer system, wherein the primary and secondary heatexchanging chambers of the stator are operably connected to the primaryand secondary heat exchanging circuits of the system, respectively, sothat the primary and secondary heat transferring portions of the rotorstructure are in thermal communication with the same while the heattransfer engine is operated in its cooling mode;

[0052]FIG. 8B is a schematic representation of the heat transfer engineof the first illustrative embodiment of the present invention installedwithin a heat transfer system, wherein the primary and secondary heatexchanging chambers of the stator are operably connected to the primaryand secondary heat exchanging circuits of the system, respectively, sothat the primary and secondary heat transferring portions of the rotorstructure are in thermal communication with the same while the heattransfer engine is operated in its heating mode;

[0053]FIG. 9 is a graphical representation of the closed-loop operatingcharacteristic of the heat transfer engine of the present invention(i.e. with the primary and secondary heat exchanging portions of therotor in thermal communication with primary and secondary heatexchanging circuits of a heat transfer system), showing the ideal rateof heat exchange from the primary portion of the rotor to the secondaryportion thereof, as a function of angular velocity of the rotor aboutits axis of rotation;

[0054]FIGS. 10A, 10B and 10C, collectively, show a flow chartillustrating the steps of the control process carried out by thetemperature-responsive system controller of the heat transfer engine ofthe present invention, operated in either its cooling or heating mode;

[0055]FIG. 11A is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid and gaseous phases of refrigerant within the rotor structurethereof when the heat transfer engine is at rest prior to entering thecooling mode;

[0056]FIG. 11B is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid, gaseous and vapor phases of refrigerant within the rotorstructure thereof during the first few revolutions thereof during thefirst stages of start up operation in its cooling mode;

[0057]FIG. 11C is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid, homogeneous fluid, vapor and gaseous phases of refrigerantwithin the rotor structure thereof during the second stage of start upoperation in its cooling mode;

[0058]FIG. 11D is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid, homogeneous fluid, vapor and gaseous phases of refrigerantwithin the rotor structure thereof when vapor compression begins withinthe centrifugal heat transfer engine during the third stage of start upoperation in its cooling mode;

[0059]FIG. 11E is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid, homogeneous fluid, vapor and gaseous phases of refrigerantwithin the rotor structure thereof during the fourth stage of start-upoperation in its cooling mode;

[0060]FIG. 11F is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid, homogeneous fluid, vapor and gaseous phases of refrigerantwithin the of the rotor structure of the heat transfer engine of FIG. 1rotor structure thereof as vapor compression occurs during the fifthstage of start-up operation in its cooling mode;

[0061]FIG. 11G is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid, homogeneous fluid, vapor and gaseous phases of refrigerantwithin the rotor structure as superdeheating and condensation beginduring the sixth stage of start-up operation in its cooling mode;

[0062]FIG. 11H is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid, homogeneous fluid, vapor and gaseous phases of refrigerantwithin the rotor structure thereof during the seventh stage of start upoperation in its cooling mode;

[0063]FIG. 11I is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid, homogeneous fluid, vapor and gaseous phases of refrigerantwithin the rotor structure during the eight (i.e. steady-state) stage ofoperation in its cooling mode;

[0064]FIG. 12A is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid and gaseous phases of refrigerant within the rotor structurethereof when the centrifugal heat transfer engine is at rest prior toentering its heating mode;

[0065]FIG. 12B is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid, gaseous and vapor phases of refrigerant within the rotorstructure thereof during the first few revolutions thereof during thefirst stages of start up operation in its heating mode;

[0066]FIG. 12C is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid, homogeneous fluid, vapor and gaseous phases of refrigerantwithin the rotor structure thereof during the second stage of start upoperation in its heating mode;

[0067]FIG. 12D is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid, homogeneous fluid, vapor and gaseous phases of refrigerantwithin the rotor structure when vapor compression begins within thecentrifugal heat transfer engine during the third stage of start upoperation in the heating mode;

[0068]FIG. 12E is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid, homogeneous fluid, vapor and gaseous phases of refrigerantwithin the rotor structure thereof during the fourth stage of start-upoperation in its heating mode;

[0069]FIG. 12F is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid, homogeneous fluid, vapor and gaseous phases of refrigerantwithin the rotor structure thereof as vapor compression occurs duringthe fifth stage of start-up operation in its heating mode;

[0070]FIG. 12G is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid, homogeneous fluid, vapor and gaseous phases of refrigerantwithin the rotor structure as superdeheating and condensation beginduring the sixth stage of start-up operation in its heating mode;

[0071]FIG. 12H is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid, homogeneous fluid, vapor and gaseous phases of refrigerantwithin the rotor structure thereof during the seventh stage of start upoperation in the heating mode;

[0072]FIG. 12I is a schematic representation of the rotor structure ofthe heat transfer engine of FIG. 1, showing the physical location of theliquid, homogeneous fluid, vapor and gaseous phases of refrigerantwithin the rotor structure thereof during the eight (i.e. steady-state)stage of operation in the heating mode;

[0073]FIG. 13 is an elevated, partially cut-away view of a roof-mountedair-conditioning system, in which the centrifugal heat transfer engineof the first illustrative embodiment is integrated with conventional airreturn and supply ducts that extend into and out of structuralcomponents of a building;

[0074]FIG. 14A is an elevated cross-sectional view of the centrifugalheat transfer engine of the second illustrative embodiment of thepresent invention, showing its fluid-carrying rotor structure rotatablysupported in a precasted stator housing having primary and secondaryfluid input and outport ports connectable to primary and secondary heatexchanging circuits, respectively, so that heat exchanging fluidcyclically flowing therethrough passes over a multiplicity of turbineblades affixed to the rotor structure and imparts torque thereto inorder to maintain the angular velocity thereof in accordance with itstemperature-responsive controller;

[0075]FIG. 14B is an elevated end view of the centrifugal heat transferengine of FIG. 14A, showing flanged fluid conduit connections forconnection to primary and secondary heat exchanging circuits;

[0076]FIG. 15A is an elevated transparent side view of the rotorstructure of the heat transfer engine shown in FIGS. 14A and 14B,removed from its stator housing, showing spiraled geometric similaritiesbetween the primary and secondary heat transfer portions of the heattransfer engine of first illustrative embodiment shown in FIG. 1 and theprimary and secondary heat transfer portions of the heat transfer engineof the second illustrative embodiment shown in FIG. 14A and 14B;

[0077]FIG. 15B is an elevated exploded view of the fluid-circulatingrotor structure of the second illustrative embodiment shown in FIGS. 14Aand 14B, removed from its stator housing, showing how the precastedrotor disc structures are joined together to provide an integralstructure within which a self-circulating closed fluid circuit is formedand how the suction shaft screw and throttling device orifice areinserted into the rotor shaft assembly;

[0078]FIG. 15C is an elevated side view of the spiraled suction screwand throttling device orifice of the rotor structure of the heattransfer engine of the second illustrative embodiment;

[0079]FIG. 15D is a side view of the threaded port cap and gasket beingfitted on the charging end of the rotor structure of the heat transferengine of the second illustrative embodiment of the present invention;

[0080]FIG. 15E is an elevated end view of a vaned rotor disk of thesecond illustrative embodiment, showing a spiraled portion of the fluidcarrying circuit formed therein and the turbine vane slots machined inthe surfaces thereof;

[0081]FIG. 15F is two elevated views of a turbine vane of the heattransfer engine of the second illustrative embodiment, showing the vanebase and illustrating a possible blade surface configuration;

[0082]FIG. 15G is an elevated side view of a vaned rotor disc of therotor of the heat transfer engine of FIGS. 14A and 14B, showing itsturbine vanes, and a machined fluid passageway portion formed in therotor structure thereof;

[0083]FIG. 15H is an elevated end view of the first end rotor disk ofthe secondary heat transfer portion of the rotor shown in FIG. 15B,showing its spiraled portion of the fluid carrying circuit formedtherein;

[0084]FIG. 15I is an elevated, side view of the first rotor end disc ofthe secondary heat transfer portion of the rotor shown in FIG. 15B;

[0085]FIG. 15J is an elevated end view of the first rotor end disc ofthe primary heat transfer portion of the rotor of FIG. 15B, showing itsspiraled portion of the fluid carrying circuit formed therein;

[0086]FIG. 15K is an elevated side view of the first rotor end disc ofthe primary heat transfer portion of the rotor of FIG. 15B, showing itsspiraled portion of the fluid carrying circuit formed therein;

[0087]FIG. 15L is an elevated transparent side view of thefluid-carrying rotor structure of the second illustrative embodiment ofthe heat transfer engine hereof, shown removed from the stator portionthereof with the closed fluid carrying circuit embedded within a heatconductive, solid-body rotor structure;

[0088]FIG. 16A is a schematic representation of the rotor structure ofthe heat transfer engine of FIGS. 14A and 14B, showing the physicallocation of the liquid and gaseous phases of refrigerant within therotor structure thereof when the heat transfer engine hereof is at restprior to entering its cooling mode;

[0089]FIG. 16B is a schematic representation of the rotor structure ofthe heat transfer engine of FIGS. 14A and 14B, showing the physicallocation of the liquid, gaseous and vapor phases of refrigerant withinthe rotor structure during the first few revolutions thereof during thefirst stages of start up operation in the cooling mode;

[0090]FIG. 16C is a schematic representation of the rotor structure ofthe heat transfer engine of FIGS. 14A and 14B, showing the physicallocation of the liquid, homogeneous fluid, vapor and gaseous phases ofrefrigerant within the rotor structure during the second stage of startup operation in the cooling mode;

[0091]FIG. 16D is a schematic representation of the rotor structure ofthe heat transfer engine of FIGS. 14A and 14B, showing the physicallocation of the liquid, homogeneous fluid, vapor and gaseous phases ofrefrigerant within the rotor structure when vapor compression beginswithin the heat transfer engine during the third stage of start upoperation in its cooling mode;

[0092]FIG. 16E is a schematic representation of the rotor structure ofthe heat transfer engine of FIGS. 14A and 14B, showing the physicallocation of the liquid, homogeneous fluid, vapor and gaseous phases ofrefrigerant within the rotor structure during the fourth stage ofstart-up operation in its cooling mode;

[0093]FIG. 16F is a schematic representation of the rotor structure ofthe heat transfer engine of FIGS. 14A and 14B, showing the physicallocation of the liquid, homogeneous fluid, vapor and gaseous phases ofrefrigerant within the rotor structure as superdeheating andcondensation begin during the sixth stage of start-up operation in itscooling mode;

[0094]FIG. 16G is a schematic representation of the rotor structure ofthe heat transfer engine of FIGS. 14A and 14B, showing the physicallocation of the liquid, homogeneous fluid, vapor and gaseous phases ofrefrigerant within the rotor structure during the seventh andsteady-state state of start up operation in its cooling mode;

[0095]FIG. 16H is a schematic representation of the rotor structure ofthe heat transfer engine of FIGS. 14A and 14B, showing the physicallocation of the liquid, homogeneous fluid, vapor and gaseous phases ofrefrigerant within the rotor structure during the eighth state stage ofoperation, at an angular velocity exceeding steady-state, in its coolingmode;

[0096]FIG. 17 is a schematic diagram of a heat transfer system, in whichthe heat transfer engine of the second illustrative embodiment isarranged so that the rotor structure thereof is rotated by fluid (water)flowing through the secondary heat exchanging fluid circuit, while theangular velocity thereof is controlled using a pump and flow controlvalve controlled by the temperature-responsive system controller;

[0097]FIG. 18 is a schematic diagram of a heat transfer system, in whicha turbine-based heat transfer engine of the present invention isarranged so that the rotor structure thereof is rotated by an electricmotor in direct connection with the rotor, while water from a coolingtower is circulated through the primary heat exchanging circuit;

[0098]FIG. 19 is a schematic diagram of a heat transfer system, in whichthe primary heat exchanging chamber of a first turbine-based centrifugalheat transfer engine hereof is connected to the secondary heatexchanging chamber of a second turbine-like heat transfer engine hereof,whereas the primary heat transfer chamber of the secondary turbine-likeheat transfer engine is in fluid communication with a cooling towerwhile the secondary heat exchanging chamber of the second turbine-likeheat transfer engine is in fluid communication with fluid supplycircuit;

[0099]FIG. 20 is a schematic diagram of a hybrid heat transfer engine,in which the primary heat transfer portion of the rotor is realized ascoiled structure mounted on a common shaft and contained within aprimary heat transfer chamber of the coiled heat transfer engine of thefirst illustrative embodiment, whereas the secondary heat transferportion of the rotor is realized as a turbine-like finned structuremounted on the common shaft and contained with a secondary heat transferchamber of the turbine-like heat transfer engine of the secondillustrative embodiment, shown operated in its cooling mode;

[0100]FIG. 21 is a schematic diagram of the hybrid heat transfer engineof FIG. 20, wherein the primary heat transfer portion thereof functionsas an air or gas conditioning evaporator while the secondary heattransfer portion functions as a condenser in an open loop fluid cooledcondenser, driven by an electric motor connected directly to the rotorshaft by way of a magnetic torque converter;

[0101]FIG. 22 is a schematic diagram of a heat transfer system of thepresent invention embodied within an automobile, wherein the rotor ofthe heat transfer engine is rotated by an electric motor driven byelectrical power supplied through a power control circuit, and producedby the automobile battery recharged by an alternator within the enginecompartment;

[0102]FIG. 23 is a schematic diagram of a heat transfer system of thepresent invention embodied within an refrigerated tractor trailer truck,wherein the rotor of the heat transfer engine is rotated by an electricmotor driven by electrical power supplied through a power controlcircuit and produced by a bank of batteries recharged by an alternatorwithin the engine compartment;

[0103]FIG. 24 is a schematic diagram of a heat transfer system of thepresent invention embodied within an aircraft equipped with a pluralityof heat transfer engines of the present invention, wherein the rotor ofeach heat transfer engine is rotated by an electric motor driven byelectrical power supplied through voltage regulator and temperaturecontrol circuit, and produced by an onboard electric generator;

[0104]FIG. 25 is a schematic diagram of a heat transfer system of thepresent invention embodied within a refrigerated freight train equippedwith a plurality of heat transfer engines of the present invention,wherein the rotor of each heat transfer engine is rotated by an electricmotor driven by electrical power supplied through voltage regulator andtemperature control circuit, and produced by an onboard pneumaticallydriven electric generator; and

[0105]FIG. 26 is a schematic diagram of a heat transfer system of thepresent invention embodied within a refrigerated shipping vesselequipped with a plurality of heat transfer engines of the presentinvention, wherein the rotor of each heat transfer engine is rotated byan electric motor driven by electrical power supplied through voltageregulator and temperature control circuit, and produced by an onboardpneumatically driven electric generator.

DETAILED DESCRIPTION OF THE ILLUSTRATIVE EMBODIMENTS OF THE PRESENTINVENTION

[0106] Referring to the Figures of the accompanying Drawings, theIllustrative Embodiments of the Present Invention will be described ingreat detail below. Throughout the drawings, like structures will berepresented by like reference numerals.

[0107] First Illustrative Embodiment Of The Heat Transfer Engine Hereof

[0108] In FIG. 1, a first illustrative embodiment of the centrifugalheat transfer engine is shown. As shown, this embodiment of the heattransfer engine comprises a rotatable structure (i.e. “rotor”) realizedas a spiral coiled tubing assembly, that is rotatably supported by astationary structure (“stator”). Thus, hereinafter this embodiment ofthe heat transfer engine shall be referred to as the coiled centrifugalheat transfer engine.

[0109] As shown in FIG. 1, reversible centrifugal heat transfer engine 1comprises a number of major system components, namely: a stator housing2; primary port connection assembly 3; secondary port connectionassembly 4; heat-exchanging rotor 5; a heat carrying medium 6; torquegenerator 7; temperature selection unit 9; temperature sensors 9Athrough 9D; primary and secondary fluid flow rate controllers 10A and10B; and temperature-responsive system controller 11. Each of thesesystem components will be described in detail below.

[0110] As shown, the stator housing comprises primary and secondary heattransfer chambers 13 and 14, and a thermal isolation barrier 15 disposedtherebetween. By definition, the primary heat transfer chamber shallindicate hereinafter and in the claims the environment within which thetemperature of a fluid (i.e. gas or liquid) contained therein is to bemaintained by way of operation of the heat transfer engine hereof.Primary heat transfer chamber 13 has inlet and outlet ports 16A and 16B,and secondary heat transfer chamber 14 has inlet and outlet ports 16Cand 16D. Primary port connection assembly 3 is provided forinterconnecting a primary heat exchanging circuit 20 (e.g. ductwork) tothe inlet and outlet ports of the primary heat transfer chamber, so asto permit a primary heat exchanging medium 21, such as air or water, toflow through the primary heat exchanging circuit and the primary heatexchanging chamber during the operation of the heat transfer engine,while the primary and secondary heat exchanging circuits are insubstantial thermal isolation of each other. Similarly, secondary portconnection assembly 4 is provided for interconnecting a secondary heatexchanging circuit 22 to the inlet and outlet ports of the secondaryheat transfer chamber, so as to permit a secondary heat exchangingmedium 23 to flow through the secondary heat exchanging circuit and thesecondary heat transfer chamber during the operation of the heattransfer engine, while the primary and secondary heat exchangingcircuits are in substantial thermal isolation of each other.

[0111] As illustrated in FIG. 1, heat exchanging rotor 5 is rotatablysupported within the stator housing 2 about an axis of rotation 25 andhas a substantially symmetrical moment of inertia about the axis ofrotation. The heat exchanging rotor has a primary heat exchanging endportion 2A disposed within the primary heat transfer chamber 13, asecondary heat exchanging end portion 2B disposed within the secondaryheat transfer chamber 14, and an intermediate portion 2C disposedbetween the primary and secondary heat exchanging end portions 2A and2B. As shown in FIGS. 2A and 2B, the heat exchanging rotor 5 contains aclosed fluid circuit 32 symmetrically arranged about the axis ofrotation and has a return portion 26A extending along direction of theaxis of rotation. The primary heat exchanging end portion 2A of therotor is disposed in thermal communication with the primary heatexchanging circuit 20, whereas the secondary heat exchanging end portion2B of the rotor is disposed in thermal communication with the secondaryheat exchanging circuit 22. The intermediate portion 2C thereof isphysically adjacent to the thermal isolation barrier 15. The physicalarrangement described above presents substantially high thermalresistance to heat transfer between the primary and secondary heatexchanging chambers 13 and 14 during operation of the reversible heattransfer engine.

[0112] As shown in FIG. 1, stator structure 2 is realized as a pair ofrotor support elements 27A and 27B mounted upon a support platform 28 ina spaced apart manner.

[0113] In the illustrative embodiment, a predetermined amount of a heatcarrying medium 6, such as refrigerant, is contained within the closedfluid circuit 32 and 26A of the rotor. In general, the heat carryingmedium is characterized by three basic thermodynamic properties: (i) itspredetermined heat of evaporation at which the heat carrying mediumtransforms from liquid phase to vapor phase; and (ii) its predeterminedheat of condensation at which the heat carrying medium transforms fromvapor phase to liquid phase; and (iii) direction reversibility of phasechange of the heat carrying liquid. Examples of suitable refrigerantsfor use with the heat transfer engine hereof include fluid refrigerantshaving a liquid or gaseous state during applicable operating temperatureand pressure ranges. When selecting a refrigerant, the followingconsideration should be made: compatibility between the refrigerant andmaterials used to construct the closed fluid flow passageway; chemicalstability of the refrigerant under conditions of use; applicable safetycodes (e.g. non-flammable refrigerants made be required); toxicity; costfactors; and availability.

[0114] In accordance with the principles of the present invention, therefrigerant or other heat-exchanging medium contained within the closedfluid circulation circuit 32 is self-circulating, in that it flowscyclically throughout the closed fluid circulation circuit in responseto rotation of the heat exchanging rotor. By virtue of the geometry ofthe closed fluid circulation circuit about the rotational axis of therotor, a complex distribution of centrifugal forces act upon and causethe contained refrigerant to circulate within the closed fluidcirculation circuit in a cyclical manner, without the use of externalpumps or other external fluid pressure generating devices. Conceivably,there exist a family of geometries for the closed fluid circulationcircuit which, when embodied within the rotor, will generate asufficient distribution of centrifugal forces to cause self-circulationof the contained fluid in response to rotation of the rotor. However,the double spiral-coil geometry with the spiral return path along therotor central axis has been discovered to be the preferred geometry ofthe present invention. Thus, in each of the three major embodiments ofthe rotor structure of the present invention, the double spiral coilgeometry is shown embodied in a rotor structure of one form or another.

[0115] The function of the torque generator 7 is to impart torque to theheat exchanging rotor 5 in order to rotate the same about its axis ofrotation at a predetermined angular velocity. In general, the torquegenerator may be realized in a variety of ways using known technology.Electric, hydraulic and pneumatic motors are just a few types of torquegenerators that may be coupled to the rotor shaft 29 and be used tocontrollably impart torque thereto under the control of systemcontroller 11.

[0116] The function of the temperature selecting unit 9 is to select(i.e. set) a temperature which is to be maintained along at least aportion of the primary heat exchanging circuit 20. In the illustrativeembodiment, the temperature selecting unit 9 is realized by electroniccircuitry having memory for storing a selected temperature value, andmeans for producing an electrical signal representative thereof. Thetemperature sensors 9A, 9B, 9C, and 9D located at inlet and outlet ports16A, 16B, 16C and 16D may be realized using any state of the arttemperature sensing technology. The function of such devices is tomeasure the temperature of the primary heat exchanging medium 21 flowingthrough the inlet and outlet ports of the primary heat exchangingchamber 13, and the secondary heat exchanging medium 23 flowing throughthe inlet and outlet ports of the secondary heat exchanging chamber 14,and produce electrical signals representative thereof for use by thesystem controller 11 as will be described in greater detail hereinafter.

[0117] The function of the primary and secondary fluid flow ratecontrollers 10A and 10B is to control the rate of flow of primary andsecondary heat exchanging fluid within the primary and secondary heatexchanging circuits, respectively. In other words, the function of theprimary fluid flow rate controller 10A is to control the rate of heatflow between the primary heat exchanging portion of the rotor and theprimary heat exchanging circuit passing through the primary heatexchanging chamber of the stator housing. Similarly, the function of thesecondary fluid flow rate controller 10B is to control the rate of heatflow between the secondary heat exchanging portion of the rotor and thesecondary heat exchanging circuit passing through the secondary heatexchanging chamber of the stator housing. In the illustrativeembodiments, the fluid flow rate controllers are controlled by thetemperature responsive system controller 11 of the engine.

[0118] Primary and secondary fluid flow rate controller 10A and 10B maybe realized in a variety of ways depending on the nature of the heatexchanging medium being circulated through primary and secondary heatexchanging chambers 13 and 14 as the rotor is rotatably supported withinthe stator. For example, when the primary heat exchanging medium is airported from the environment in which the air temperature is to bemaintained, then primary fluid flow controller 10A may be realized by anair flow control valve (e.g. damper), whose aperture dimensions areelectromechanically controlled by electrical control signals produced bythe system controller. When the primary heat exchanging medium is waterported from a primary heat exchanging circuit in which the watertemperature is to be maintained, then primary fluid flow controller maybe realized by an water control flow valve, whose aperture dimensionsare electromechanically controlled by electrical control signalsproduced by the system controller. In either case, the function of theprimary fluid flow rate controller is to control the flow rate of theprimary heat exchanging medium flowing through the primary heatexchanging chamber in response to the sensed temperature of the heatexchanging medium at either the inlet or outlet port in either theprimary or secondary heat exchanging chambers, and the temperatureselected by temperature selection unit. Greater details with regard tothis aspect of the control process will be described hereafter.

[0119] The secondary fluid flow rate controller 10B may be realized in amanner similar to the primary fluid flow rate controller 10A. In fact,it is possible to construct a heat transfer engine in which the primaryand secondary heat exchange fluids are different in physical state (e.g.the primary heat exchange fluid can be air, while the secondary heatexchange fluid is water, and vice versa). In each possible case, thefunction of the secondary fluid flow rate controller is to control theflow rate of the secondary heat exchanging medium flowing through thesecondary heat exchanging chamber, in response to the sensed temperatureof the heat exchanging medium at either the inlet or outlet port ineither the primary or secondary heat exchanging chambers and thetemperature selected by temperature selection unit.

[0120] The system controller 11 of the present invention has severalother functions, namely: to read the temperature of the ambientoperating environment measured by way of temperature sensors 9, 9A, 9B,9C, and 9D; and in response thereto, generate suitable control signalswhich directly control the operation of torque generator 7; andindirectly control the angular velocity of the heat exchanging rotor,relative to the stator; and control the fluid flow rate of the primaryand secondary heat exchanging fluids 21 and 23 flowing through theprimary and secondary heat exchanging chambers 13 and 14, respectively.The need to control the angular velocity of the heat exchanging rotor,and the flow rates of the primary and secondary heat exchanging fluidswill be described in detail hereinafter with reference to thethermodynamic refrigeration process of the present invention.

[0121] In general, the reversible heat transfer engine of the presentinvention has two modes of operation, namely: a heating mode which isrealized when the heat exchanging rotor is rotated in a firstpredetermined direction of rotation; and a cooling mode which isrealized when the rotor is rotated in a second predetermined directionof rotation. Also, while it would be desired that the enclosure (i.e.stator) of the system be thermally insulated for optimal heat transferoperation and efficiency, this is not an essential requirement forsystem operation.

[0122] Referring to FIGS. 2A through 7, the structure and functions ofthe heat exchanging rotor of the first illustrative embodiment will nowbe described in greater detail below. As shown, heat exchanging rotor 5of the first illustrative embodiment is realized as a length of tubing32 symmetrically coiled around support shaft 29 extending along the axisof rotation of the rotor. As shown, the tubing assembly 36 and 37 has adouble spiral-coil geometry, and the support shaft contains a spiralreturn passage 33 formed therethrough with an inlet opening 34 and anoutlet opening 35. The spiral-coiled tubing assembly has a first spiraltubing portion 36, a second spiral tubing portion and bidirectionalmetering device 38 disposed therebetween. As shown, the ends of thefirst and second spiral tubing portions 36 and 37 are attached to boththe inlet 52 and outlet 53 openings of the spiral return passage 33along the rotor shaft and creates the closed fluid circulation circuitwithin the heat transfer structure. The function of the bi-directionalmetering device 38 is to control (1) the rate of flow of liquidrefrigerant into the second spiral tubing portion 36 and (2) the amountof pressure drop between the secondary and primary tubing portionsduring a preselected range of rotor angular velocities (RPM). Theoptimum rotor angular velocity is arrived at and controlled by thesystem controller in response to temperature changes in the air orliquid being treated by the heat transfer engine of the presentinvention. The reason the throttling device 38 is bi-directional is toallow for refrigerant flow reversal when the direction of rotor rotationis reversed when switching from the cooling mode to the heating mode ofthe heat transfer engine.

[0123] By virtue of the geometry of the closed fluid circulation circuit26 realized within the rotor, a complex distribution of centrifugalforces are generated and act upon the molecules of refrigerant containedwithin the closed circuit in response to rotation of the rotor relativeto its stator. This, in turn, causes refrigerant to cyclically circulatewithin the closed circuit, without the use of external pumps or otherexternal fluid pressure generating devices.

[0124] In FIGS. 4A and 4B, details relating to the construction of rotorshaft 29 of the first illustrative embodiment are shown. In particular,the rotor shaft 29 comprises a central shaft core 40 of solidconstruction enclosed within as cylindrical tube cover 41. Also, acharging port 42 is provided along the end of the central tube in orderto provide access to refrigerant inside the closed (i.e. sealed)self-circulating fluid circulation circuit (i.e. system). As best shownin FIG. 4A, central shaft core 40 has a spiraled passage 33 formed aboutthe outer surface thereof, and is enclosed within tube cover 41, therebycreating a spiral shaped passageway 33 from one end of the rotor shaftto the other end thereof. As shown in FIGS. 5, 6A and 6B, a pair ofholes 44 are drilled through cylindrical tube cover 41 into the spiraledpassageway 33 at the ends of the central shaft 29A and 29B. These holesallow the first and second end portions of double-coil tubing assemblyto interconnect with the ends of the spiral rotor shaft, and thus formthe closed fluid circulation circuit within the rotor structure.

[0125] As shown in FIGS. 7A, 7B and 7C, the rotor of the firstillustrative embodiment also includes a plurality of tubing supportbrackets 45A, 45B, 45C and 45D for support of the spiraled tubularsections thereof in position about its central shaft. As shown, each ofthese tubing support brackets comprises shaft attachment means 45extending from the rotor shaft 29, and tubing support element 46 forsupporting a selected portion of the tubing assembly spiraled about therotor shaft. These tubing support brackets may be made from any suitablematerial such as metal, composite material, or other functionallyequivalent material. In general, the tubing used to realize the rotor ofthe first illustrative embodiment may vary in inner diameter as thediameter of the tubing around the central shaft varies. Preferably, theexterior surface of the rotor tubing is finned, while the internalsurface thereof is rifled as this construction will improve the heattransfer function of the rotor.

[0126] Having described the structure and function of the systemcomponents of the heat transfer engine of the first illustrativeembodiment, it is appropriate at this juncture to describe in greaterdetail the operation of the system controller in each of the heattransfer modes of operation of the engine.

[0127] In FIG. 10A, the heat transfer engine hereof is shown installedin an environment 50 through which the primary heat exchanging circuit20 passes in order to control the temperature thereof while the engineis operated in its cooling mode. While the medium within thisillustrative environment will typically be ambient air, it is understoodthat other mediums may be temperature maintained in differentapplications. Notably, in FIG. 10A, the closed fluid flow circuit ofrotor is arranged according to the first conFiguration. To specify thedirection of rotor shaft rotation in this mode of operation, it ishelpful to embed a Cartesian Coordinate system in the stator, so thatthe +z axis and point of origin thereof are aligned with the +z axis andpoint of origin of the rotor. In the first rotor conFiguration, thedirection of the rotor rotation is counterclockwise about the +z axis ofthe stator reference system when the engine is operated in its coolingmode.

[0128] In FIG. 10B, the heat transfer engine hereof is shown installedin the same environment 50 shown in FIG. 10B, while the engine isoperated in its heating mode. In FIG. 10B, the closed fluid flow circuitof rotor is arranged once again according to the first rotorconFiguration. To specify the direction of rotor shaft rotation in thismode of operation, it is helpful to embed a Cartesian Coordinate systemin the stator, so that the +z axis and point of origin thereof arealigned with the +z axis and point of origin of the rotor. In the firstrotor configuration, the direction of the rotor rotation is clockwiseabout the +z axis of the stator reference system when the engine isoperated in its heating mode.

[0129] In FIGS. 18 and 19, an alternative embodiment of the heatexchanging rotor is schematically illustrated. As shown, the rotor 52 isrealized as a solid body having first and second end portions 2A and 2Bof truncated cone-like geometry, connected by a central cylindricalportion 2C extending about an axis of rotation. As illustrated, a closedfluid flow circuit 26 having essentially the same geometry as rotor 5 ofthe first illustrative embodiment is embodied (or embedded) within thesolid rotor body. As such, this embodiment shall be referred to as theembedded rotor embodiment of the present invention. As in the firstillustrative embodiment, the closed fluid circuit of rotor 52symmetrically extends about its rotor axis of rotation. Alsobi-directional metering device 38 is realized within the central portionof the rotor body, as shown. Preferably, one end of the rotor has anaccess port 95 and 96, (e.g. a removable screw cap) for introducingrefrigerant into or removing refrigerant from the closed fluid flowcircuit. The fluid flow circuit may be realized in the solid body of therotor in a variety of ways. One way is to produce a solid rotor body intwo symmetrical half sections using injection molding techniques, sothat respective portions of the closed fluid flow circuit are integrallyformed therein. Thereafter, the molded body halves can be joinedtogether using appropriate gaskets, seals and fastening techniques.Advanced composite materials, including ceramics, may be used toconstruct the rotor body. Alternatively, as shown in FIGS. 15A to 15K,the rotor may be realized by assembling a plurality of rotor discs, eachembodying a portion of the closed fluid flow circuit. Details regardingthis alternative embodiment will be described in greater detailhereinafter.

[0130] In order to properly construct the rotor, the direction ofrotation of the spiral tubing along the closed fluid flow circuit isessential. To specify this tubing direction, it is helpful to specifythe portion of the fluid flow circuit along the rotor shaft (i.e. therotor axis) as the inner fluid flow path, and the portion of the fluidflow circuit extending outside of the rotor shaft as the outer fluidflow path. Notably, the outer fluid flow path is bisected by thebi-directional metering device into a first outer fluid flow pathportion and a second outer fluid flow path portion. The end section ofthese outer fluid flow path portions away from the metering deviceconnect with the end sections of the inner fluid flow path, to completethe closed fluid flow path within the heat exchanging rotor. In order tospecify the direction of spiral of the above-defined fluid flow pathportions, it is helpful to embed a Cartesian Coordinate system withinthe rotor such that the point of origin of the reference system islocated at one end of the rotor shaft and the +z axis of the referencesystem extends along the axis of rotation (i.e shaft) of the rotortowards the other end of the shaft. With the reference system installed,there are two possible ways of configuring the closed fluid flow circuitof the rotor of the present invention.

[0131] According to the first possible conFiguration, looking from thepoint of origin of the reference system down the +z axis, the firstouter fluid flow portion extends spirally about the +z axis incounter-clockwise (CCW) direction from the first end portion of theshaft to the metering device, and then continues to extend spirallyabout the +z axis in a counter-clockwise (CCW) from the metering deviceto the second end portion of the rotor shaft; and looking from the pointof origin of the reference system down the +z axis, the inner fluid flowpath extends spirally about the +z axis in a clockwise(CW) direction.

[0132] According to the second possible conFiguration, as shown in FIGS.14A, 14B, 18, and 19, looking from the point of origin of the referencesystem down the +z axis, the first outer fluid flow portion extendsspirally about the +z axis in a counter-clockwise (CCW) direction fromthe first end portion 26 of the shaft to the inlet of the fluid flowtube 84 as shown in FIG. 17A, and then continues to extend spirallyabout the +z axis in counter-clockwise (CCW) from the fluid flow tubedevice to the second end portion of the rotor shaft; looking from thepoint of origin of the reference system down the +z axis, the innerfluid flow path extends spirally about the +z axis in acounter-clockwise direction (CCW). Either of these two conFigurationswill work in a functionally equivalent manner. However, as will bedescribed in greater detail below, depending on the rotor configurationemployed in any particular application, the direction of shaft rotationwill be different for each heat transfer mode (e.g. cooling mode orheating mode) selected by the system user.

[0133] Principles Of Throttling Device Design

[0134] It will be helpful to now describe some practical principleswhich can be used to design and construct the throttling (i.e. metering)device within the rotor structure hereof.

[0135] In general, the function of the throttling device of the presentinvention is to assist in the transformation of liquid refrigerant intovapor refrigerant without impacting the function of the rotor within theheat transfer engine hereof. In general, this system component (i.e. themetering device) is realized by providing a fluid flow passagewaybetween the condenser functioning portion of the rotor and theevaporator functioning portion. This fluid flow passageway has an innercross-sectional area that is smaller than the smallest innercross-sectional area of the evaporator section of the rotor. Inprinciple, there are many different ways to realize the reducedcross-sectional area in the fluid flow passageway between the primaryand secondary heat exchanging sections of the rotor. Regardless of howthis system component is realized, a properly designed metering devicewill operate in a bi-directional manner (i.e., in the cooling or heatingmode of operation). The function of the metering device is to providethe necessary pressure drop between the condensor and evaporatorfunctioning portions of the heat transfer engine hereof, and allowsufficient Superheat to be generated across the evaporator functioningportion of the rotor. In the case of the illustrative embodiments, themetering device should be designed to provide optimum fluid flowcharacteristics between the primary and secondary heat transfer portionsof the rotor.

[0136] For example, in the first illustrative embodiment where theprimary and secondary heat exchanging portions are made from hollowtubing of substantially equal diameter, the metering device can beeasily realized by welding (or brazing) a section of hollow tubingbetween the primary and secondary heat exchanging portions, having aninner diameter smaller than the inner diameter of the primary andsecondary heat exchanging portions. In order to provide optimum fluidflow characteristics across the metering device, the ends of the smallreduced diameter tubing section can be flared so that the inner diameterof this small tubing section is matched to the inner diameter of thetubing from which the primary and secondary heat exchanging portions aremade. In an alternative embodiment, it is conceivable that tubing of theprimary and secondary heat exchanging portions can be continuouslyconnected by welding or brazing process and that the metering device canbe realized by crimping or stretching the tubing adjacent to theconnection, to achieve the necessary reduction in fluid flow passageway.

[0137] In the second illustrative embodiment disclosed herein, theclosed fluid passageway is realized within a solid-body rotor structuresuitable for turbine type application where various types of fluid areused to input torque to the rotor during engine operation. In thisparticular embodiment, the metering device can be easily realized bywelding (or brazing) a section of hollow tubing between the primary andsecondary heat exchanging portions, having an inner diameter smallerthan the inner diameter of the primary and secondary heat exchangingportions, as shown in FIG. 18.

[0138] In yet another alternative embodiment, a plurality of meteringdevices of the type described above can be used in parallel in order toachieve the necessary reduction in fluid flow passageway, and thus asufficient pressure drop thereacross the primary and secondary heatexchanging portions of the rotor. In such an alternative embodiment, itis understood that the condenser functioning portion of the rotor wouldterminate in a first manifold-like structure, to which the individualmetering devices would be attached at one end. Similarly, the evaporatorportion of the rotor would terminate in a second manifold-likestructure, to which the individual metering devices would be attached attheir other end.

[0139] In any particular embodiment of the rotor of the presentinvention, it will be necessary to design and construct the meteringdevice so that system performance parameters are satisfied. In thepreferred embodiment, a reiterative design procedure is used to designand construct the metering device so that system performancespecifications are satisfied by the operative engine construction. Thisdesign and construction procedure will be described below.

[0140] The first step of the design method involves determining thesystem design parameters which include, for example: the ThermalTransfer Capacity of the system measured in BTUs/hour; Thermal Load onthe system measured in BTUs/hour; the physical dimensions of the rotor;and volume and type of refrigerant contained within the rotor (less than80% of internal volume). The second step involves specifying the designparameters for the metering device which, as described above, includeprimarily the smallest cross-sectional area of the fluid passagewaybetween the first and second heat exchanging portion of the rotor.According to the method of the present invention, it is not necessary tocalculate the metering device design parameters using a thermodynamic orother type of mathematical model. Rather, according to the method of thepresent invention, an initial value for the metering device designparameters (i.e. the smallest cross-sectional area of the fluidpassageway) is selected and used to construct a metering device forinstallation within the rotor structure of the system under design.

[0141] The next step of the design method involves attaching infra-redtemperature sensors to the inlet and outlet ports of theevaporator-functioning portion of the rotor, and then connecting thesetemperature sensors to an electronic (i.e. computer-based) recordinginstrument well known in the temperature instrumentation art. Then,after (i) constructing the heat transfer engine according to thespecified system design parameters, (ii) loading refrigerant into therotor structure, and (iii) setting the primary design parameter (i.e.,smallest cross-sectional area) in the metering device, the heat transferengine is operated under the specified thermal loading conditions forwhich it was designed. When steady-state operation is attained,temperature measurements at the inlet and outlet ports of the rotorevaporator, T_(ei) and T_(eo), respectively, are taken and recordedusing the above-described instrument. These measurements are then usedto determine whether or not the metering device produces enough of apressure drop between the condensor and evaporator so that sufficientSuperheat is produced across the evaporator to drive the engine to thedesired level of performance specified by the system design/performanceparameters described above.

[0142] This condition is detected using the following design criteria.If T_(eo) is not greater than T_(ei) by 6 degrees, then there is notenough Superheat being generated at the evaporator, or the angularvelocity of the rotor is too low. If this condition exists, then therotor angular velocity is increased to Wmax and recheck T_(ei) andT_(ei). Then if T_(eo) is not greater than T_(ei) by 6 degrees, then thesmallest cross-sectional area (e.g. diameter) through the meteringdevice is too large and a reduction therein is needed. If this conditionis detected, then the engine is stopped. The metering device is modifiedby reducing the cross-sectional area of the metering device by anincremental amount. The modified engine is then restarted and T_(ei) andT_(eo) remeasured to determine whether the amount of the Superheatproduced across the evaporator is adequate. Thereafter, the reiterativedesign process of the present invention is repeated in the mannerdescribed above until the desired amount of Superheat is produced withinthe rotor of the production prototype under design. When this conditionis achieved, the design parameters of the metering device are carefullymeasured and recorded, and the metering device at which this operatingcondition is achieved is used to design and construct “productionmodels” of the heat transfer engine. Notably, only the design model ofthe heat transfer engine requires infra-red temperature sensors forSuperheat monitoring purposes.

[0143] System Control Process of the Present Invention

[0144] Referring now to FIGS. 8A, 8B, and 10A to 10C, thetemperature-response control process of the present invention will bedescribed for both the cooling and heating modes of the centrifugal heattransfer engine.

[0145] When the rotor of the first con Figuration is rotatably supportedwithin the stator housing and rotated in the counter-clockwise directionas shown in FIG. 8A, a complex distribution of centrifugal forces areautomatically generated and act upon the molecules of refrigerantcontained within the closed circuit. This causes the refrigerant toautomatically circulate within the closed circuit in a cyclical mannerfrom the first end portion of the rotor, to the second end portionthereof, and then back to the first end portion along the spiral fluidflow path of the support shaft. In this case, the engine is operated inits cooling mode, and the spiral tubing section 36A of the rotor withinthe primary heat exchanging chamber functions as an evaporator while thespiral tubing section 37A within the secondary heat exchanging chamberfunctions as a condenser. The overall function of the rotor in thecooling mode is to transfer heat from the primary heat exchangingchamber to the secondary heat exchanging chamber under the control ofthe system controller.

[0146] When the direction of the rotor is reversed as shown in FIG. 8B,the refrigerant contained within the closed fluid circuit automaticallycirculates therewithin in a cyclical manner from the second end portionof the rotor, to the first end portion thereof, and then back to thesecond end portion along the spiral fluid flow path of the supportshaft. In this case, the engine is operated in its heating mode, and thespiral tubing section of the rotor within the primary heat exchangingchamber 36A functions as a condenser, while the spiral tubing section37A within the secondary heat exchanging chamber functions as anevaporator. The overall function of the rotor in the heating mode is totransfer heat from the secondary heat exchanging chamber to the primaryheat exchanging chamber under the control of the system controller.

[0147] In either of the above-described modes of operation, the fluidvelocity of the refrigerant within the rotor is functionally dependentupon a number of factors including, but not limited to, the angularvelocity of the rotor relative to the stator, the thermal loading uponthe first and second end portions of the rotor, and internal losses dueto surface friction of the refrigerant within the closed fluid circuits.It should also be emphasized that design factors such as the number ofspiral coils, the heat transfer quality of materials used in theirconstruction, the diameter of the spiral coils, the primary heattransfer surface area, the secondary heat transfer surface area, and therotor angular velocity, and horsepower can be varied to alter the heattransfer capacity and efficiency of the centrifugal heat transferengine.

[0148] In order to cool the ambient environment (or fluid) to theselected temperature set by thermostat 9, the heat exchanging rotor musttransfer, at a sufficient flow rate, heat from the primary heatexchanging chamber to the secondary heat exchanging chamber, from whichit can then be liberated to the secondary heat exchanging circuit andthus maintain the selected temperature in a controlled manner.Similarly, to heat the ambient environment (or fluid) to the selectedtemperature set by the thermostat, the heat exchanging rotor musttransfer, at a sufficient flow rate, heat from the secondary heatexchanging chamber to the primary heat exchanging chamber, from which itcan then be liberated to the primary heat exchanging circuit andmaintain the selected temperature in a controlled manner.

[0149] As shown in FIGS. 8A and 8B, each of the ports in the primary orsecondary heat exchanging chambers of the heat transfer engine hasinstalled within its flowpath a temperature sensor 9A through 9Doperably connected to the temperature-responsive system controller 11.The function of each of these port-located temperature sensors is tomeasure the temperature of the liquid flowing through its associatedfluid inlet or outlet port as it passes over and/or through the endportions of the rotor. Within the environment or fluid being heated,cooled or otherwise conditioned, thermostat 9 or a like control deviceprovides a means for setting a threshold or target temperature that isto be maintained within the primary heat exchanging chamber as theprimary and secondary heat exchanging fluids are caused to circulatewithin the primary and secondary heat exchanging chambers, respectively.

[0150] The primary function of the system controller is to manage theload-reduction operating characteristics of the heat transfer engine. Inthe illustrative embodiments, this is achieved by controlling (1) theangular velocity of the rotor within prespecified limits during systemoperation, and (2) the flow rate of the primary and secondary heatexchange fluids circulating through the primary and secondary heatexchange chambers of the engine, respectively. As will be describedbelow in connection with the control process of FIGS. 10A to 10C,rotor-velocity and fluid flow-rate control is achieved by maintainingparticular port-temperature constraints (i.e. conditions) on a real-timebasis during the operation of the system in its designated mode ofoperation. In the illustrative embodiment of the present invention,these temperature constraints are expressed as difference equationswhich establish constraints (i.e. relations) among particular sensedtemperature parameters.

[0151] As illustrated, on the chart shown in FIG. 9; as the rotor RPMω^(L) increases upward from zero to a point of intersection betweenω_(L) and Q_(L), the following conditions exist: (1) Load controlbegins; (2) the spiraled return passageway is clear of liquidrefrigerant; (3) about two thirds of the primary heat transfer portionis occupied by liquid refrigerant; (4) the secondary heat transferportion is about 85 percent of fully occupied by liquid refrigerant; (5)all flow control devices are within 10 percent of maximum flow. Thesystem controller 11, gradually, continues to increase the RPM ω up toω_(H). Control over the quantity of heat transferred Q is maintainedbetween Q_(L) (low load) and Q_(H) (high load). The temperature controldifferential is ΔQ, (ΔQ=Q_(H)-Q_(L)), and the range of temperaturecontrol selected on the temperature selector 9 is limited by the designcapacity of the particular heat transfer engine at hand. As shown inFIG. 9, if the RPM ω exceeds ω_(H), the refrigeration effect begins todecrease for one of two reasons: (1) the load has diminished to a pointwhere no heat is available to be transferred in functional quantities;and (2) the weight of the liquid refrigerant in the liquidpressurization length by centrifugal forces exceeds pressurizing forcesexerted on the refrigerant by the liquid pressurization lengths spiraledstructure. Optimum operating conditions for the heat transfer engine arebetween ω_(L) and ω_(H), and Q_(L) and Q_(H). The intersectionsindicated are dictated by thermal capacity, refrigerant type and volume,and application, and are located by operational calibration.

[0152] As illustrated in FIGS. 10A to 10C, these temperature constraintsof the system control process are maintained by the system controllerduring cooling or heating modes, respectively. These temperatureconstraints depend on the ambient reference temperature T1 set bythermostat 9, and the temperatures sensed at each port of the first andsecondary heat exchanging circuits of the system. The process by whichthe system controller controls the rotor velocity and fluid flow ratesin the primary and secondary heat exchanging chambers will be describedin detail below.

[0153] In FIGS. 10A to 10C, the system control program of theillustrative embodiment is shown in the form of a computer flow diagram.During the operation of the heat transfer engine, the system controllerexecutes the control program in a cyclical manner in order toautomatically control the rotor velocity and fluid flow rates withinprespecified operating conditions, while achieving the desired degree oftemperature control along the primary heat exchanging circuit. Duringexecution of the control process, the plurality of data storageregisters associated with the system controller 11 are periodically readby its microprocessor. Each of these data storage registers isperiodically (e.g. 10 times per second) provided with a new digital wordproduced from its respective A/D converter associated with thetemperature sensor (9A, 9B, 9C, 9D) measuring the sensed temperaturevalue. Thus during the execution of the control program, the datastorage registers associated with the system controller are updated withcurrent temperature values measured at the input and output ports of theprimary and secondary heat exchanging chambers of the system.

[0154] As indicated at Block A in FIG. 10A, the first step of thecontrol process involves initializing all of the temperature dataregisters of the system. Then at Block B the microprocessor reads thecode (i.e. data) from the temperature data registers and then at Block Cthe Mode Selection Control determines whether the cooling or heatingmode has been selected by the user. If the cooling mode has beenselected at Block C, then the system controller enters Block D andcontrols the torque generator (e.g. motor) so that the rotor is rotatedin the CCW direction up to about 10% of the maximum design velocityω_(H), while the primary and secondary fluid flow rate controllers arecontrolled to allow fluid flow rates up to about 10 percent (10%) of themaximum flow rate. At Block E, the angular velocity of the rotor iscontrolled by the microprocessor performing the following rotor-velocitycontrol operations represented by the following rules: ifΔT₁=T_(a)−T_(t)≧2° F., then increase rotor velocity ω at rate of onepercent per minute up to ω_(H); and if ΔT₁=T_(a)−T_(t)≦2° F., thenreduce the rotor-velocity ω at a rate of one percent per minute down toω_(L).

[0155] At Block F, the primary fluid flow rate is controlled by themicroprocessor by performing the following primary fluid-flow ratecontrol operations: if ΔT₁=T_(a)−T_(t)≧2° F., and ΔT₁=T_(a)−T_(t)≧10°F., then increase the fluid flow rate of the primary heat exchangingfluid by one percent per minute up to PFRmax; and if ΔT₁=T_(a)−T_(t)≦0°F., then reduce the fluid flow rate of the primary heat exchanging fluidby one percent per minute down to PFRmin.

[0156] Notably, an increase in the rate of primary heat exchanging fluidthrough the primary heat exchanging chamber affects the refrigerationcycle by increasing the rate and amount of heat flowing from the primaryheat transfer portion of the rotor to the secondary heat transferportion thereof, as illustrated by the heat transfer loop in FIG. 8A. Asthe temperature of the primary heat transfer portion of the rotorincreases due to an increase in the heat exchange fluid flow (PFR), morerefrigerant is evaporated (i.e. boiled off) and more of the primary heattransfer portion is occupied by vapor. Consequently, more of thesecondary heat transfer portion of the rotor is occupied by liquidrefrigerant and the increased liquid pressurization length causes theBubble Point within the closed fluid flow circuit to move furtherdownstream along the throttling device length (closer to the evaporatorfunctioning section).

[0157] At Block G, the secondary fluid flow rate is controlled by themicroprocessor by performing the following secondary fluid-flow ratecontrol operations: ΔT₃=T_(d)−T_(c)≧2° F., or, ΔT₃=T_(d)−T_(c)≧40° F.,and ΔT₁=T_(a)−T_(t)≧2° F., then increase the fluid flow rate of thesecondary heat exchanging fluid by one percent per minute up to SFRmax;and if ΔT₃=T_(d)−T_(c)≧20° F., or ΔT₁=T_(c)−T_(t)≦2° F., then reduce thefluid flow rate of the primary heat exchanging fluid by one percent perminute down to SFRmin.

[0158] After performing the operations at Blocks E, F and G, themicroprocessor reads once again the temperature values in itstemperature value storage registers, and then at Block J determineswhether there has been any change in mode (e.g. switch from the coolingmode to the heating mode). If no change in mode has been detected atBlock J, then the microprocessor reenters the control loop defined byBlocks E through H and performs the operations specified therein tocontrol the angular velocity of the rotor to and the flow rates of theprimary and secondary fluid flow-rate controllers, PFR and SFR

[0159] If at Block J in FIG. 10B the microprocessor determines whetherthe mode of the heat transfer engine has been changed (e.g. from thecooling mode to the heating mode) then the microprocessor returns toBlock C in FIG. 10A and then proceeds to Block K. At Block K themicroprocessor controls the torque generator (e.g. motor) so that therotor is rotated in the CW direction up to about 10% of the maximumdesign velocity ω_(H), while the primary and secondary fluid flow ratecontrollers are controlled to allow fluid flow rates up to about 10percent (10%) of the maximum flow rate. At Block L, the angular velocityof the rotor is controlled by the microprocessor performing thefollowing rotor-velocity control operations: if ΔT₄=T_(t)−T_(a)≧2° F.,then increase rotor velocity ω at a rate of one percent per minute up toω_(H); and if ΔT₄=T_(a)−T_(t)≧20° F., then reduce the rotor-velocity Cat a rate of one percent per minute down to ω_(L).

[0160] At Block M, the primary fluid flow rate is controlled by themicroprocessor by performing the following primary fluid-flow ratecontrol operations: if ΔT₄=T_(t)−T_(a)≧20° F., and ΔT₅=T_(b)−T_(a)≧20°F., then increase the fluid flow rate of the primary heat exchangingfluid by one percent per minute up to PFRmax; and if ΔT₄=T_(t)−T_(a)≦2°F., then reduce the fluid flow rate of the primary heat exchanging fluidby one percent per minute down to SFRmax.

[0161] Notably, an increase in the rate of secondary heat exchangingfluid through the secondary heat exchanging chamber affects therefrigeration cycle by increasing the rate and amount of heat flowingfrom the secondary heat transfer portion of the rotor to the primaryheat transfer portion thereof, as illustrated by the heat transfer loopin FIG. 8B. As the temperature of the secondary heat transfer portion ofthe rotor increases because of a heat exchange fluid flow increase(SFR), more refrigerant is evaporated (i.e. boiled off) and more of thesecondary heat transfer portion of the rotor is occupied by vapor.Consequently, more of the primary heat transfer portion of the rotor isoccupied by liquid refrigerant and the increased Liquid PressurizationLength causes the Bubble Point to move further upstream along thethrottling device length of the (closer to the secondary heat transferportion of the rotor).

[0162] At Block N, the secondary fluid flow rate is controlled by themicroprocessor by performing the following secondary fluid-flow ratecontrol operations: if ΔT₅=T_(c)−T_(d)≧10° F., or ΔT₅=T_(c)−T_(d)≦40°F., and ΔT₄=T_(t)−T_(c)≧2° F., then increase the fluid flow rate of thesecondary heat exchanging fluid by one percent per minute up to SFRmax;and if ΔT₅=T_(c)−T_(d)≧20° F., then reduce the fluid flow rate of theprimary heat exchanging fluid by one percent per minute down to SFRmin.

[0163] After performing the operations at Blocks L, M and N, themicroprocessor reads once again the temperature values in thetemperature value storage register of the system controller, and atBlock P determines whether there has been any change in mode (e.g.switch from heating mode to cooling mode). If no change in mode has beendetected at Block P, then the microcontroller reenters the control loopdefined by Blocks L through N and performs such operations in order tocontrol the angular velocity of the rotor and the flow rates of theprimary and secondary fluid flow-rate controllers. If at Block P in FIG.10C the microprocessor determines that the mode of the heat transferengine has been changed (e.g. from the heating mode to the cooling mode)then the microprocessor returns to Block C in FIG. 10A and then proceedsto Block D. Notably, the speed at which the microprocessor traversesthrough this control loops described above will typically besubstantially greater than the rate at which the temperature values maychange as indicated by the data values in the temperature storageregisters. Thus the system controller can easily track thethermodynamics of the heat transfer engine of the present invention.

[0164] In the illustrative embodiment, the parameters (Wmax, Wmin,PFRmax, PRFmin, SFRmax, SFRmin) employed in the control processdescribed above may be determined in a variety of ways.

[0165] In the illustrative embodiment, the parameters (W_(H), W_(L),PFRmax, PFRmin, SFRmax, and SFRmin) employed in the control processdescribed above may be determined in a variety of ways. W_(H) (rotorRPM) is primarily determined by the strength of materials used toconstruct the rotor, and, secondly, at an RPM where Q_(H) is realized.Q_(H) is found by acquiring the temperature of the fluid entering theprimary heat transfer portion and the temperature of the fluid leavingthe primary heat transfer portion. The lowest of the two temperature issubtracted from the highest temperature and the sum is the fluidtemperature difference. The fluid temperature difference multiplied bythe specific heat of the fluid being used equals the BTU per pound thatparticular fluid has absorbed or dissipated. W_(L) is determined whenthe RPM is reduced to a point where no appreciable net refrigerationaffect is taking place. PFRmax can be gallons per minute (GPM) forliquids or cubic feet per minute (CFM) for gasses. For example, waterentering the primary heat transfer portion at a temperature of 60° F.and leaving the primary heat transfer portion at 50° F. has atemperature difference of 10° F. Water has a specific heat of 1 BTU perpound at temperatures between 32° F. and 212° F. Therefore, waterrecirculated at 100 gallons per minute, having a temperature differenceof 10° F. is transferring 60,000 BTU per hour. Five tons ofrefrigeration and 60,000 BTUH heating. Air entering the primary heattransfer portion at a temperature of 60° F. and leaving the primary heattransfer at 50° F. has a temperature difference of 10° F. and contains22 BTU per pound (dry air and associated moisture). Air at 60° F. and 50percent relative humidity also contains approximately 22 BTU per pound(dry air and associated moisture). The Sensible Heat Ratio(SHR=Q_(s)/Q_(t)) is arrived at by dividing the quantity of sensibleheat in the air (Q_(s)) by the total amount of heat in the air (Q_(t)).The sensible heat ratio of the 60° F. air in the above example is 0.46and the sensible heat ratio of the 50° F. air is 0.73. The 60° F. aircontains mostly latent heat, about 11.88 BTU latent heat and 10.12 BTUsensible heat. The 50° F. air contains most sensible heat, about 5.94BTU latent heat and 16.06 BTU sensible heat. The net refrigerationaffect is the difference between 11.88 BTU and 5.94 BTU, or 5.94 BTU perpound of recirculated air has been transferred from the air into theprimary heat transfer portion. In that condition, the air contains 13.01cubic feet of air per pound. The air contracts slightly during cooling,about 0.19 cubic foot per pound of dry air, and, if 2,000 cubic feet ofair are recirculated per minute, the net refrigeration affect will be544,788.24 BTU per hour, or 4.57 tons of refrigeration. In this example,PFRmax would be 2000 CFM and SFRmax will equal PFRmax because of thelack of heat being introduced into the self-circulating circuit frominternal motor windings and the heat of compression caused byreciprocating compressors. The range between PFRmin and PFRmax, andSFRmin and SFRmax is determined by the physical aspects of a particularinstallation. Physical aspects can range from total environmental loadreduction control system to a simple on-off control circuit.

[0166] Referring to FIGS. 11A to 11I, the refrigeration process of thepresent invention will now be described with the heat transfer engine ofthe present engine being operated in its cooling mode of operation.Notably, each of these drawings schematically depicts, from across-sectional perspective, both the first and second heat exchangingportions of the rotor. This presentation of the internal structure ofthe closed fluid passageway throughout the rotor provides a clearillustration of both the location and the state of the refrigerant alongthe closed fluid passageway thereof.

[0167] As shown in FIG. 11A, the rotor is shown at its rest position,which is indicated by the absence of any rotational arrow about therotor shaft. At this stage of operation, the internal volume of theclosed fluid circuit is occupied by about 65% of refrigerant in itsliquid state. Notably, the entire spiral return passageway along therotor shaft is occupied with liquid refrigerant, while the heatexchanging portions of the rotor are occupied with liquid refrigerant ata level set by gravity in the normal course. The portion of the fluidpassageway above the liquid level in the rotor is occupied byrefrigerant in a gaseous state. The closed fluid passageway isthoroughly cleaned and dehydrated prior to the addition of the selectedrefrigerant to prevent any contamination thereof.

[0168] As shown in FIG. 11B, the rotor is rotated in a counter-clockwise(CCW) direction within the stator housing of the heat transfer engine.During steady state operation in the cooling mode, illustrated in FIGS.11G to 11I, the primary heat transfer portion will perform a liquidrefrigerant evaporating function, while the secondary heat transferportion performs a refrigerant vapor condensing function. However, atthe stage of operation indicated in FIG. 1B, the liquid refrigerantwithin the spiraled passageway of the shaft begins to flow into thesecondary heat transfer (i.e. exchanging) portion of the rotor andoccupies the entire volume thereof. As shown, a very small portion (i.e.about one coil turn) of the primary heat transfer portion is occupied byrefrigerant vapor as it passes through the throttling (i.e. metering )device, while the remainder of the primary heat transfer portion of therotor and a portion of the spiraled passageway of the shaft onceoccupied by liquid refrigerant is occupied with gas. Notably, theboundary between the length of liquid refrigerant and length of gas (orrefrigerant vapor) in the rotor is, by definition, the “Liquid Seal” andresides along the primary heat transfer portion of the rotor shaft atthis early stage of start-up operation. In general, the Liquid Seal islocated between the condensation and throttling processes supportedwithin the rotor. The Liquid Seal has two primary functions within therotor, namely: during start-up operations, to occlude the passage ofrefrigerant vapor, thereby forcing the vapor to condense in thesecondary heat transfer portion (i.e. condenser); and, more precisely,during steady state operation the Liquid Seal resides at a point alongthe length of the secondary heat transfer portion where enoughrefrigerant vapor has condensed into a liquid by absorbing “LatentHeat”, thereby occupying the total internal face area of the passageway.As used hereinafter, the term “Latent Heat” is defined herein as theheat absorbed by (into) the liquid refrigerant (homogeneous fluid)during the evaporation process, as well as the heat discharged from thegaseous refrigerant during the condensation process.

[0169] Liquid refrigerant contained in the first one half of thesecondary heat transfer portion between the rotor shaft and the point ofhighest radius (from the center of rotation) is effectively moved andpartially pressurized by centrifugal force, and the physical shape ofthe spiraled passageway, outwardly from the center of rotation into thesecond one half of the secondary heat transfer portion. Liquidrefrigerant contained in the second one half of the secondary heattransfer portion between the point of highest radius (from the center ofrotation) and the throttling device (i.e. metering) is effectivelypressurized (against flow restriction caused by the throttling deviceand Liquid Seal) by the physical shape of the spiraled passageway andcentrifugal force. This section of the secondary heat transfer portionof the rotor which varies in response to “Thermal Loading” is definedherein as the “Liquid Pressurization Length”. The term “Thermal Load” or“Thermal Loading” as used here shall mean the demand of heat transferimposed upon the heat transfer engine of the present invention in aparticular mode of operation. Liquid refrigerant is pressurized due to(i) the distribution of centrifugal forces acting on the molecules ofthe liquid refrigerant therein as well as (ii) the pressure created bythe liquid refrigerant being forcibly driven into the secondary heattransfer portion against the Liquid Seal Hand the metering device flowrestriction.

[0170] As shown in FIG. 11B, during the start up stage of engineoperation in a counterclockwise (CCW) direction, the Liquid Seal movestowards the secondary heat transfer portion, land refrigerant flowinginto the primary heat transfer portion is restricted by the throttlingdevice sand the refrigerant stacks up in the secondary heat transferportion. Very little refrigerant flows into the primary heat transferportion, and no refrigeration affect has yet taken place. The smallamount of vapor in the primary heat transfer portion will gather some“Superheat” which will remain in the vapor and gaseous refrigerantwithin the primary heat transfer portion, as a result of the LiquidSeal. As will be used hereinafter, the term “Superheat” shall be definedas a sensible heat gain above the saturation temperature of the liquidrefrigerant, at which a change in temperature of the refrigerant gasoccurs (sensed) with no change in pressure.

[0171] As shown in FIG. 11C, the rotor continues to increase in speed inthe CCW direction. At this stage of operation, the Liquid PressurizationLength of the refrigerant begins to create enough pressure within thesecondary heat transfer portion to overcome the pressure restrictioncaused by the throttling device and thus liquid begins to flow into theprimary heat transfer portion of the rotor. As shown, the Liquid Sealhas moved along the rotor shaft towards the secondary heat transferportion.

[0172] At this stage of operation, refrigerant beyond the meteringdevice and into about the first spiral coil of the primary heat transferportion is in the form of a “homogeneous fluid” (i.e. a mixture ofliquid and vapor state) while a portion of the first spiral coil and aportion of the second one contain refrigerant in its homogeneous state.As used hereinafter, the term “homogeneous fluid” shall mean a mixtureof flash gas and low temperature, low pressure, liquid refrigerantexperiencing a change-in-state (the process of evaporation) due to itsabsorption of heat. The length of refrigerant over which Evaporationoccurs shall be defined as the Evaporation Length of the refrigerant,whereas the section of the refrigerant stream along the fluid flowpassageway containing gas shall be defined as the Superheat Length, asshown. The homogeneous fluid entering the primary heat transfer portion“displaces” the gas therewithin, thereby pushing it downstream into thespiraled passageway of the rotor shaft. Throttling of liquid refrigerantinto vapor absorbs heat from the primary heat transfer portion of therotor, imparting “Superheat” to the gaseous refrigerant. A “cooler”vapor created by the process of throttling enters the primary heattransfer portion and begins to absorb more Superheat. Refrigerant gasand vapor are compressed between the homogeneous fluid in the primaryheat transfer portion and the Liquid Seal in the spiraled passageway ofthe rotor shaft.

[0173] Notably, at this stage of operation shown in FIG. 11C, there isonly enough pressure in the Secondary heat transfer section to cause aminimal amount of liquid to flow into the primary heat transfer portionof the rotor, and thus throttling (i.e. partially evaporating) occursslightly. Consequently, the refrigeration affect has begun slightly andthe only heat being absorbed by the refrigerant is Superheat in theSuperheat Length of the refrigerant stream. The vapor beginning formjust downstream in the primary heat transfer portion is “Flash” gas fromthe throttling process.

[0174] The stage of operation represented in FIG. 11C illustrates whatshall be called the “Liquid Line”. As shown, the Liquid Line shall bedefined as the point where the homogeneous fluid ends and the vaporbegins along the length of the primary heat transfer portion. Therefore,the liquid line illustrated in FIGS. 11C to 11F can occupy a shortlength of the primary heat transfer portion as a mixture of homogeneousfluid and a very dense vapor which extends downstream to the Superheatlength. The exact location along the primary heat transfer portion willvary depending on the quantity of homogeneous fluid, which is inproportion to the amount of heat being absorbed and the Thermal Load(i.e. heat transfer demand) being imposed on the heat transfer engine inits mode of operation. The Liquid Line is not to be confused with theLiquid Seal.

[0175] As the rotor continues to increase to its steady state speed inthe CCW direction, as shown in FIG. 11D, the amount of refrigerant vaporin the primary heat transfer portion increases due to increasedthrottling and increased “Flash” gas entering the same. The effect ofthis is to increase the quantity of homogeneous fluid entering theprimary heat transfer portion of the rotor. As shown in FIG. 11D, theLiquid Seal has moved even further along the rotor shaft towards thesecondary heat transfer portion. Also, less liquid refrigerant occupiesthe spiraled passageway of the rotor shaft, while more homogeneous fluidoccupies the primary heat transfer portion of the rotor (i.e. in theform of Superheat). Also as indicated, the direction of heat flow isfrom the primary heat transfer portion to the secondary heat transferportion. However at this stage of operation, this heat flow is trappedbehind the Liquid Seal in the spiraled passageway of the shaft.

[0176] As the rotor continues to increase to its steady state speed inthe CCW direction, as shown in FIG. 11E, the quantity of refrigerantvapor within the primary heat transfer portion of the rotor continues toincrease due to the increased production of flash gas from thethrottling of liquid refrigerant. As shown, the Liquid Seal has movedtowards the end of the rotor shaft and the secondary heat transferportion inlet thereof. Also, during this stage of operation, the flow ofheat (i.e. Superheat) from the primary heat transfer portion is stilltrapped behind the Liquid Seal in the spiraled passageway of the rotorshaft. Consequently, the Superheat Heat from the primary heat transferportion is unable to pass onto the secondary heat transfer portionsprimary and secondary heat transfer surfaces, and thus optimal operationis not yet achieved at this stage of engine operation. During this stageof operation some heat (Superheat) may transfer into the rotor shaftfrom the refrigerant vapor if the shaft temperature is less that thetemperature of the refrigerant vapor; and some heat may transfer intothe refrigerant vapor if the refrigerant vapor temperature is less thanthat of the rotor shaft. The rotor shaft and its internal spiraledpassageway is a systematic source of primary and secondary Superheattransfer surfaces where heat can be either introduced into the vapor ordischarged from the vapor. Heat caused by rotor shaft bearing frictionis absorbed by the refrigerant vapor along the length of the rotor shaftand can add to the amount of Superheat entering the secondary heattransfer portion. This additional Superheat further increases thetemperature difference between the Superheated vapor and the secondaryheat transfer surfaces of the secondary heat transfer portion which, inturn, increases the rate of heat flow from the Superheated vapor within.Consequently, this enhances necessary heat transfer locations needed toachieve steady state operation.

[0177] At the stage of operation shown in FIG. 11F, the rotor isapproaching its steady-state angular velocity, and is shown operating inthe CCW direction of operation at what shall be called “ThresholdVelocity”. As shown, the remaining liquid refrigerant in the rotor shaftis now completely displaced by refrigerant vapor produced as a result ofthe evaporation of the liquid refrigerant in a primary heat transferportion of the rotor. Consequently, Superheat produced from the primaryheat transfer portion is permitted to flow through the spiraledpassageway of the rotor shaft and into the secondary heat transferportion, where it can be liberated by way of condensation across thesecondary heat transfer portion. As shown, Superheat Length of therefrigerant stream within the primary heat transfer portion of the rotorhas decreased, while the evaporation length of the refrigerant streamhas increased proportionally, indicating that the refrigeration effectwithin the primary heat transfer portion is increasing.

[0178] At the stage of operation shown in FIG. 11F, the Liquid Seal isno longer located along the rotor shaft, but within the secondary heattransfer portion of the rotor, near the end of the rotor shaft. Vaporcompression begins to occur in the last part of the primary heattransfer portion and along the spiraled passageway of the rotor. At thisstage of operation the pressure of the liquid refrigerant in the LiquidPressurization Length has increased sufficiently enough to furtherincrease the production of homogeneous fluid in the primary heattransfer portion. This also causes the quantity of liquid in thesecondary heat transfer portion to decrease “Pulling” on the flash gasand vapor located in the spiraled passageway in the rotor shaft, and inthe primary heat transfer portion downstream from the homogeneous fluid.The pulling affect enhances vapor compression taking place in thespiraled passageway in the rotor shaft. At this stage of operation thehomogeneous fluid is evaporating absorbing heat within the primary heattransfer portion of the rotor for transference and systematic dischargefrom the secondary heat transfer portion. In other words, during thisstage of operation, the vapor within the primary heat transfer portioncan contain more Superheat by volume than the gas with which it ismixed. Thus, the increased volume in dense vapor in the primary heattransfer portion provides a means of storing Superheat (absorbed fromthe primary heat exchanging circuit) until the vapor stream flows intothe secondary heat transfer portion of the rotor where it can beliberated to the secondary heat exchanging circuit by way of conduction.

[0179] As shown in FIG. 11G, the heat transfer engine of the presentinvention is operated at what shall be called the “Balance PointCondition”, the refrigeration cycle of which is illustrated in FIG. 17Aand 17B. At this stage of operation, the refrigerant within the rotorhas attained the necessary phase distribution where simultaneously thereis an equal amount of refrigerant being evaporated in the primary heattransfer portion as there is refrigerant vapor being condensed in thesecondary heat transfer portion of the rotor.

[0180] As shown in FIG. 11G, the Superheat that has “accumulated” in therefrigerant vapor during the start up sequence shown in FIGS. 11Athrough 11F begins to dissipate from the DeSuperheat Length of therefrigerant stream along the secondary heat transfer portion of therotor. The density of the refrigerant gas increases, and vaporcompression occurs as the Superheat is carried by the refrigerant gasfrom the Superheat Length of the primary heat transfer portion to theDeSuperheat Length in the secondary heat transfer portion by thespiraled passageway in the rotor shaft. Thus, as the Superheat isdissipated in the secondary heat transfer portion and compressed vaporin the secondary heat transfer portion begins to condense into liquidrefrigerant, a denser vapor remains. Consequently, the spiraledpassageway of the rotor shaft has a greater compressive affect on thevapor therein at this stage of operation. In other words, the spiraledpassageway of the shaft is pressurizing the Superheated gas and densevapor against the Liquid Seal in the secondary heat transfer portion.

[0181] As shown in FIG. 11G, pressurization of liquid refrigerant in thesecondary heat transfer portion of the rotor pushes the liquidrefrigerant through the throttling device at a higher pressure,sufficiently enough, which causes a portion of the liquid refrigerant to“flash” into a gas, thereby, reducing the temperature of the remaininghomogeneous fluid (i.e. liquid and dense vapor) entering the primaryheat transfer portion thereof. The liquid refrigerant portion of thehomogeneous fluid, in turn, evaporates, creating sufficient vaporpressure therein that it displaces vapor downstream within the primaryheat transfer portion into the spiraled passageway of the rotor shaft.This vapor pressure, enhanced by vapor compression caused by thespiraled passageway in the rotor shaft, pushes the same into thesecondary heat transfer portion of the rotor, where its Superheat isliberated over the DeSuperheat Length thereof.

[0182] At the Balance Point condition, a number of conditions existthroughout steady-state operation. Foremost, the Liquid Seal tends toremain near the same location in the secondary heat transfer portion,while the Liquid Line tends to remain near the same location in theprimary heat transfer portion. Secondly, the temperature and pressure ofthe refrigerant in the secondary heat transfer portion of the rotor ishigher than the refrigerant in the primary heat transfer portionthereof. Third, the rate of heat transfer from the primary heatexchanging chamber of the engine into the primary heat transfer portionthereof is substantially equal to the rate of heat transfer from thesecondary heat transfer portion of the engine into the secondary heatexchanging chamber thereof. Thus, if the primary heat transfer portionof the rotor is absorbing heat at about 12,000 BTUH from the primaryheat exchanging circuit, then the secondary heat transfer portionthereof is dissipating about 12,000 BTUH to the secondary heatexchanging circuit.

[0183] In order to appreciate the heat transfer process supported by theengine of the present invention, it will be helpful to focus on therefrigerant throttling process within the rotor in slightly greaterdetail.

[0184] The throttling process of the present invention can be describedin terms of the three sub-processes which determine the condition of therefrigerant as it passes through the throttling device of the engine ineither of its rotational directions. These sub-processes are defined asthe Liquid Length, the Bubble Point, and the Two Phase Length. Forpurposes of clarity, the sub-Processes of the throttling process will bedescribed as they occur during start-up operations and steady-stateoperations.

[0185] The Liquid Length begins at the inlet of the throttling deviceand continues to the Bubble Point. The Bubble Point exists at pointinside (or along) the throttling device, (i) at which the Liquid Length(liquid refrigerant) is separated or distinguishable from the Two PhaseLength (foamy, liquid and vapor refrigerant) and (ii) where enoughpressure drop along the restrictive passage of the throttling device hasoccurred to cause a portion of the liquid refrigerant to evaporate (asingle bubble) and reduce the temperature of the surrounding liquidrefrigerant (two phase, bubbles and liquid) for delivery into theevaporator section of the rotor. The Latent Heat given up by the liquidrefrigerant during its change in state at the Bubble Point is containedwithin the bubbles produced at the Bubble Point. Heat absorbed by thesebubbles in the evaporator section of the rotor is Superheat. The BubblePoint can exist anywhere along the throttling devices length dependingon the amount of thermal load imposed on the heat transfer engine. TheLiquid Length extends over that portion of the throttling devicecontaining pure liquid refrigerant up to the Bubble Point. The Two-PhaseLength extends from the Bubble Point into the evaporator inlet of therotor and (foamy, liquid and vapor refrigerant).

[0186] During optimum load conditions in the cooling mode, theCondensation Length and Evaporation Length each contain an equal amountof liquid refrigerant. This is because the amount of heat entering theprimary heat transfer portion of the rotor is equal to the amount ofheat leaving the secondary heat transfer portion thereof. During higherthan design load conditions (above optimum) in the cooling mode ofoperation, there is more liquid refrigerant in the secondary heattransfer portion of the rotor than in the primary heat transfer portionthereof. There are two reasons of explanation for this phenomenon. Thefirst reason is that the primary heat transfer portion of the rotor hasa higher rate of heat transfer by virtue of the higher-than-designtemperature difference existing between the homogeneous fluid in theprimary heat transfer portion of the rotor and the air or liquid passingover the primary heat transfer surfaces. The second reason is that theincrease in the throttling process lowers the temperature and pressureof the homogeneous fluid entering the primary heat transfer portion ofthe rotor. The additional liquid refrigerant in the secondary heattransfer portion of the rotor reduces the available internal volumeneeded for adequate vapor-to-liquid condensation. Operating under thesehigher-than-design load conditions, the centrifugal heat transfer engineis “Over Loaded”. In such cases, a larger rotor should be used for theapplication. An increase in the rotor RPM will cause a higher rate ofhomogeneous fluid to flow into the primary heat transfer portion.However, if the increase in RPM, and a consequent increase incentrifugal force upon the liquid refrigerant, causes the weight of theliquid refrigerant in the Liquid Pressurization Length (of the secondaryheat transfer portion) to overcome the coriolis affect, then therefrigeration cycle will cease.

[0187] When the design operating temperature of the heat exchangingfluid circulating through the primary heat exchanging chamber is belowfreezing, a defrost cycle can occur by reducing the RPM of the rotatablestructure, reducing the refrigeration affect.

[0188] During lower-than-design load conditions (below optimum) thecentrifugal heat transfer engine has more liquid refrigerant in theprimary heat transfer portion than is contained by the secondary heattransfer portion. The accumulation of liquid refrigerant in the primaryheat transfer portion is due to the low rate of heat transfer in theprimary heat transfer portion. The temperature and pressure of therefrigerant in the secondary heat transfer portion can be increased byreducing the rate of flow of the heat exchanging fluid circulatingthrough the secondary heat exchanging chamber. Such a decrease in fluidflow causes an increase in temperature and pressure of the refrigerantin the primary heat transfer portion which, in turn, causes an increasein temperature and pressure of the refrigerant in the primary heattransfer portion. The increase in temperature and pressure of therefrigerant in the primary heat transfer portion increases the amount ofheat (BTU) per pound that a hydrocarbon refrigerant is capable ofabsorbing, to an optimum saturation temperature and pressure. Theindustry design standard is 95 degrees Fahrenheit condensingtemperature. Such a controlled decrease in fluid flow shall be referredto as “Secondary Pressure Stabilization”. Such a controlled decrease influid flow can increase the engines coefficient of performance (COP, orBTU/WATT) of the heat transfer engine. A similar increase or decrease inthe primary heat exchanging fluid flow shall be referred to as “PrimaryPressure Stabilization”. During the cooling mode of operation, and whenthe centrifugal heat transfer engine has satisfied the loadrequirements, reaching a Set Point or Balance Point, the RPM of therotor can be reduced causing a reduction in the refrigeration affect tosatisfy a lesser load demand. This type of operation, or mode, is calledLoad Reduction Control (or Unloading). Unlike Unloading, thermal Loadingis where the rotor RPM is increased to satisfy a higher load demand.

[0189] The location of the Liquid Seal is affected by the amount of loadbeing exerted on the evaporation process. Liquid pressurization beginsat the Liquid Seal and occurs inside the spiraled condenser sectionalong the Liquid Pressurization Length up to the inlet of the throttling(i.e. metering) device inlet. Starting at the Liquid Seal, as the rotorrotates, the liquid refrigerant is forced toward the central axis ofrotation by the spiraled shape of the Liquid Pressurization Length inthe condenser functioning section of the rotor. The centrifugal forcesproduced during rotor rotation causes the liquid pressure to graduallyincrease along the Liquid Pressurization Length, providing a continuoussupply of higher pressure (condensed) liquid refrigerant to the inlet ofthe throttling device where the Liquid Length begins. In other words,during rotation centrifugal forces within the rotor increase the weightof the liquid refrigerant contained in the spiraled LiquidPressurization Length and cause the liquid refrigerant therewith topressurize against the flow restricting pressure drop produced by thefluid flow geometry of the throttling device, thereby completing therefrigeration cycle of the centrifugal heat transfer engine.

[0190] In FIG. 11H, the heat transfer engine of the present invention isshown operating just below its “optimum” (low load) operating condition,whereas in FIG. 11I, the heat transfer engine is shown operatedexcessively beyond its “optimum” operating condition. Notably, the term“optimum” operating condition used above is not to be equated with theterm “Balance Point” operating condition. Rather “optimum” operatingcondition is a point of operation where the amount of liquid refrigerantin the primary heat transfer portion is slightly higher than the amountof liquid refrigerant in the secondary heat transfer portion. Thisoperating point is considered optimum as the lower temperaturerefrigerant in the primary heat transfer portion is capable ofcontaining more heat (i.e. BTU per pound) than the higher pressure andtemperature liquid refrigerant contained in the secondary heat transferportion of the rotor. Consequently, during engine operation, the flowrate of heat exchanging fluid within the secondary heat exchangingchamber of the engine is reduced at times by the system controller, asthis increases the temperature of the secondary heat transfer portion(i.e. during the cooling mode), and thereby increasing the “rate” ofheat flow from the secondary heat transfer portion of the rotor(particularly on large capacity engines) into the secondary heatexchanging fluid circulating through the secondary heat exchangingchamber. If the thermal load on the engine is further reduced beyondthat shown in FIG. 11I, the spiraled passageway in the rotor shaftprevents a condition where the Liquid Pressurization Length is starvedof liquid refrigerant. This safety measure is provided by the fact thatat least sixty five percent of the total internal volume of the rotor isoccupied by refrigerant, and that quantities of refrigerant exceedingthe internal volume of the primary heat transfer portion and extendinginto the spiraled passageway in the rotor shaft are rapidly moved intothe secondary heat transfer portion (by way of the rotating spiraledpassageway along the rotor shaft), thereby rapidly replenishing theLiquid Pressurization Length thereof.

[0191] As shown in FIG. 11I, the Liquid Seal has moved nearer to thethrottling device, and even though the Liquid Seal is located in thesecondary heat transfer portion, the Liquid Pressurization Length isstill pressurizing the liquid refrigerant. In FIG. 11I, the heattransfer engine is shown operated at a point of operation where the“load” has diminished sufficiently to cause the liquid refrigerantwithin the rotor to “accumulate” in the primary heat transfer portionthereof. At this stage of operation, the system controller of the engineshould be reacting to a reduction in temperature in the primary heatexchanging chamber, thereby reducing the RPM of the rotor. Also, theflow rate controller associated with the primary heat exchanging chambershould be starting to reduce the flow rate of heat exchanging fluidcirculating within the secondary heat exchanging chamber. Notably, ifthe engine was operated in its “De-ice” or “Defrost” mode of operation,the rotor RPM would be further decreased in order to reduce therefrigeration affect. In turn, this would increase the “overall systempressure”, causing the ambient temperature about the primary heatexchanging portion to increase, thereby preventing the formation of ice(or accumulation of process fluid) on the primary and secondary heattransfer surfaces thereof.

[0192] Heat Transfer Process Of Present Invention: Heating Mode OfOperation

[0193] Referring to FIGS. 12A to 12I, the refrigeration process of thepresent invention will now be described with the heat transfer engine ofthe present engine being operation in its heating mode of operation.Notably, each of these drawings schematically depicts, from across-sectional perspective, both the first and second heat exchangingportions of the rotor. This presentation of the internal structure ofthe closed fluid passageway throughout the rotor provides a clearillustration of both the location and the state of the refrigerant alongthe closed fluid passageway thereof.

[0194] In FIG. 12A, the rotor is shown at its rest position, which isindicated by the absence of any rotational arrow about the rotor shaft.At this stage of operation, the internal volume of the closed fluidcircuit is occupied by about 65% of refrigerant in its liquid state.Notably, the entire spiral return passageway along the rotor shaft isoccupied with liquid refrigerant, while the heat exchanging portions ofthe rotor are occupied with liquid refrigerant at a level set by gravityin the normal course. The portion of the fluid passageway above theliquid level in the rotor is occupied by refrigerant in a gaseous state.The closed fluid flow passageway is thoroughly cleaned and dehydratedprior to the addition of the selected refrigerant to prevent anycontamination thereof.

[0195] As shown in FIG. 12B, the rotor is rotated in a clockwise (CW)direction within the stator housing of the heat transfer engine. Duringsteady state operation in the cooling mode, illustrated in FIGS. 12G to12I, the primary heat transfer portion will perform a liquid refrigerantevaporating function, while the secondary heat transfer portion performsa refrigerant vapor condensing function. However, at the stage ofoperation indicated in FIG. 12B, the liquid refrigerant within thespiraled passageway of the shaft begins to flow into the secondary heattransfer (i.e. exchanging) portion of the rotor and occupies the entirevolume thereof. As shown, a very small portion (i.e. about one coilturn) of the primary heat transfer portion is occupied by refrigerantvapor as it passes through the throttling (i.e. metering ) device, whilethe remainder of the primary heat transfer portion of the rotor and aportion of the spiraled passageway of the shaft once occupied by liquidrefrigerant is occupied with gas. During steady state operation theLiquid Seal resides at a point along the length of the secondary heattransfer portion where enough refrigerant vapor has condensed into aliquid thereby occupying the total internal face area of the passageway.

[0196] During the start up stage of engine operation shown in FIG. 12B,the Liquid Seal moves towards the secondary heat transfer portion, andrefrigerant flow into the primary heat transfer portion is restricted bythe throttling device and the refrigerant stacks up in the secondaryheat transfer portion. Very little refrigerant flows into the primaryheat transfer portion, and no refrigeration affect has yet taken place.The small amount of vapor in the primary heat transfer portion willgather some Superheat which will remain in the vapor and gaseousrefrigerant within the primary heat transfer portion, as a result of theLiquid Seal.

[0197] As shown in FIG. 12C, the rotor continues to increase in speed inthe CW direction. At this stage of operation, the Liquid PressurizationLength of the refrigerant begins to create enough pressure within thesecondary heat transfer portion to overcome the pressure restrictioncaused by the throttling device and thus liquid begins to flow into theprimary heat transfer portion of the rotor. As shown, the Liquid Sealhas moved along the rotor shaft towards the secondary heat transferportion. The homogeneous fluid entering the primary heat transferportion “displaces” the gas therewithin, thereby pushing it downstreaminto the spiraled passageway of the rotor shaft. Some throttling ofliquid refrigerant into vapor occurs causing enough temperature drop inthe primary heat transfer portion of the rotor and thus causing transferof Superheat into the gaseous refrigerant. A “cooler” vapor created bythe process of throttling enters the primary heat transfer portion andbegins to absorb more Superheat. Refrigerant gas and vapor arecompressed between the homogeneous fluid in the primary heat transferportion and the Liquid Seal in the spiraled passageway of the rotorshaft.

[0198] At the stage of operation shown in FIG. 12C, there is only enoughpressure in the secondary heat transfer section to cause a minimalamount of liquid to flow into the primary heat transfer portion of therotor, and therefore throttling (i.e. partially evaporating) occursslightly. Consequently, the refrigeration affect has begun slightly andthe only heat being absorbed by the refrigerant is Superheat in theSuperheat Length of the refrigerant stream. There is some vaporbeginning to form just downstream in the primary heat transfer portion,which is really “Flash” gas from the throttling process. The Liquid Lineillustrated in FIGS. 12C can occupy a short length of the primary heattransfer portion as a mixture of homogeneous fluid and a very densevapor which extends downstream to the Superheat length. The exactlocation of the Liquid Line along the primary heat transfer portion willvary depending on the quantity of homogeneous fluid, which is inproportion to the amount of heat being absorbed and the load beingimposed on it.

[0199] As the rotor continues to increase to its steady state speed inthe CW direction, as shown in FIG. 12D, the amount of refrigerant vaporin the primary heat transfer portion increases due to increasedthrottling and increased “Flash” gas entering the same. The effect ofthis is an increase in the quantity of homogeneous fluid entering theprimary heat transfer portion of the rotor. As Shown in FIG. 12D, theLiquid Seal has moved even further along the rotor shaft towards thesecondary heat transfer portion. Also, less liquid refrigerant occupiesthe spiraled passageway of the rotor shaft, while more homogeneous fluidoccupies the primary heat transfer portion of the rotor. Also asindicated, the direction of heat flow is from the primary heat transferportion to the secondary heat transfer portion (i.e. in the form ofSuperheat). However at this stage of operation, this heat flow istrapped behind the Liquid Seal in the spiraled passageway of the shaft.

[0200] As the rotor continues to increase to its steady state speed inthe CW direction, as shown in FIG. 12E, the quantity of refrigerantvapor within the primary heat transfer portion of the rotor continues toincrease due to the increased production of flash gas from throttling ofliquid refrigerant. As shown, the Liquid Seal has moved towards the endof the rotor shaft and the secondary heat transfer portion inletthereof. Also, during this stage of operation, the flow of heat (i.e.Superheat) from the primary heat transfer portion is still trappedbehind the Liquid Seal in the spiraled passageway of the rotor shaft.Consequently, the Superheat from the primary heat transfer portion isunable to pass onto the secondary heat transfer portions primary andsecondary heat transfer surfaces, and thus optimal operation is not yetachieved at this stage of engine operation. During this stage ofoperation some heat (i.e. Superheat) may transfer into the rotor shaftfrom the refrigerant vapor if the shaft temperature is less that thetemperature of the refrigerant vapor; and some heat may transfer intothe refrigerant vapor if the refrigerant vapor temperature is less thanthat of the rotor shaft.

[0201] At the stage of operation shown in FIG. 12F, the rotor isapproaching its steady-state angular velocity, and is shown operating inthe CW direction of operation at its “Threshold Velocity”. As shown, theremaining liquid refrigerant in the rotor shaft is now completelydisplaced by refrigerant vapor produced as a result of the evaporationof the liquid refrigerant in primary heat transfer portion of the rotor.Consequently, Superheat produced from the primary heat transfer portionis permitted to flow through the spiraled passageway of the rotor shaftand into the secondary heat transfer portion, where it can be liberatedby way of condensation across the secondary heat transfer portion. Asshown, Superheat Length of the refrigerant stream within the primaryheat transfer portion of the rotor has decreased, while the evaporationlength of the refrigerant stream has increased proportionally,indicating that the refrigeration effect within the primary heattransfer portion is increasing.

[0202] At the stage of operation shown in FIG. 12F, the Liquid Seal isno longer located along the rotor shaft, but within the secondary heattransfer portion of the rotor, near the end of the rotor shaft. Vaporcompression begins to occur in the last part of the primary heattransfer portion and along the spiraled passageway of the rotor. At thisstage of operation the pressure of the liquid refrigerant in the LiquidPressurization Length has increased sufficiently enough to furtherincrease the production of homogeneous fluid in the primary heattransfer portion. This also causes the quantity of liquid in thesecondary heat transfer portion to decrease “Pulling” on the flash gasand vapor located in the spiraled passageway in the rotor shaft, and inthe primary heat transfer portion downstream from the homogeneous fluid.The pulling affect enhances vapor compression taking place in thespiraled passageway in the rotor shaft. At this stage of operation, thehomogeneous fluid is evaporating absorbing heat within the primary heattransfer portion of the rotor for transference and systematic dischargefrom the secondary heat transfer portion into the heat exchanging fluidcirculating through the primary heat exchanging chamber. In other words,during this stage of operation, the vapor within the primary heattransfer portion can contain more Superheat by volume than the gas withwhich it is mixed. Thus, the increased volume in dense vapor in theprimary heat transfer portion provides a means of storing Superheat(absorbed from the primary heat exchanging circuit) until the vaporstream flows into the secondary heat transfer portion of the rotor whereit can be liberated to the secondary heat exchanging circuit by way ofconduction.

[0203] As shown in FIG. 12G, the heat transfer engine of the presentinvention is operating at what shall be called the “Balance PointCondition”. At this stage of operation, the refrigerant within the rotorhas attained the necessary phase distribution where simultaneously thereis an equal amount of refrigerant being evaporated in the primary heattransfer portion as there is refrigerant vapor being condensed in thesecondary heat transfer portion of the rotor. The secondary heattransfer portion is adding heat to the primary heat transfer chamber. Asshown in FIG. 12G, the Superheat that has “accumulated” in therefrigerant vapor during the start up sequence shown in FIGS. 12Athrough 12F begins to dissipate from the DeSuperheat Length of therefrigerant stream along the secondary heat transfer portion of therotor. The density of the refrigerant gas increases, and vaporcompression occurs as the Superheat is carried by the refrigerant gasfrom the Superheat Length of the primary heat transfer portion to theDeSuperheat Length in the secondary heat transfer portion by thespiraled passageway in the rotor shaft. Thus, as the Superheat isdissipated in the secondary heat transfer portion, and compressed vaporin the secondary heat transfer portion begins to condense into liquidrefrigerant, a denser vapor remains. Consequently, at this stage ofoperation, the spiraled passageway of the rotor shaft has a greatercompressive affect on the vapor therein. In other words, the spiraledpassageway of the shaft is pressurizing the Superheated gas and densevapor against the Liquid Seal in the secondary heat transfer portion.

[0204] As shown in FIG. 12G, pressurization of liquid refrigerant in thesecondary heat transfer portion of the rotor pushes the liquidrefrigerant through the throttling device at a sufficiently higherpressure, which causes a portion of the liquid refrigerant to “flash”into a gas, thereby, reducing the temperature of the remaininghomogeneous fluid (liquid and dense vapor) entering the primary heattransfer portion thereof. The liquid refrigerant portion of thehomogeneous fluid, in turn, evaporates which creates sufficient vaporpressure therein that it displaces vapor downstream within the primaryheat transfer portion into the spiraled passageway of the rotor shaft.This vapor pressure, enhanced by vapor compression caused by thespiraled passageway in the rotor shaft, pushes the same into thesecondary heat transfer portion of the rotor, where its Superheat isliberated over the DeSuperheat Length thereof.

[0205] At the Balance Point condition, a number of conditions existthroughout steady-state operation. Foremost, the Liquid Seal tends toremain near the same location in the secondary heat transfer portion,while the Liquid Line tends to remain near the same location in theprimary heat transfer portion. Secondly, the temperature and pressure ofthe refrigerant in the secondary heat transfer portion of the rotor ishigher than the refrigerant in the primary heat transfer portionthereof. Thirdly, the rate of heat transfer to the primary heatexchanging chamber of the engine from the secondary heat transferportion thereof is substantially equal to the rate of heat transfer fromthe primary heat transfer portion of the engine into the secondary heatexchanging chamber thereof. Thus, if the primary heat transfer portionof the rotor is absorbing heat at about 12,000 BTUH from the primaryheat exchanging circuit, then the secondary heat transfer portionthereof is dissipating about 12,000 BTUH from the secondary heatexchanging circuit.

[0206] In FIG. 12H, the heat transfer engine of the present invention isshown operating just below its optimum (low load) operating condition.In FIG. 12I, the heat transfer engine is shown operated excessivelybeyond its “optimum” operating condition. In this state, the Liquid Sealis located in the secondary heat transfer portion, and even though theLiquid Seal has moved nearer toward the throttling device, the LiquidPressurization Length is still pressurizing the liquid refrigerant. Thedemand for heat by the system controller during this state of operationhas diminished sufficiently to cause the liquid refrigerant within therotor to “accumulate” in the primary heat transfer portion thereof. Atthis stage of operation, the system controller of the engine should bereacting to an increase in temperature in the primary heat exchangingchamber, reducing the RPM of the rotor, and the flow rate controllerassociated with the primary heat transfer chamber should be starting toreduce the flow rate of the heat exchanging fluid circulating within thesecondary heat exchanging chamber.

[0207] Applications Of First Embodiment Of Heat Transfer Engine Hereof

[0208] In FIG. 13, the heat transfer engine of the first illustrativeembodiment is shown installed on the roof of a building or similarstructure, as part of an air handling system which is commonly known inthe industry as a Roof-Top or Self-Contained air conditioning unit, orair handler. In this application, the heat transfer engine functions asa roof-top air conditioning unit which can be operated in its coolingmode or heating mode. The term “air conditioning” as used herein shallinclude the concept of cooling and/or heating of the air to be“temperature conditioned”, in addition to the conditioning of air forhuman occupancy which includes its temperature, humidity, quantity, andcleanliness. As shown, the air handling unit comprises an air supplyduct 60 and an air return duct 61, both penetrating structuralcomponents of a building. The rotor of the centrifugal heat transferengine is rotated by a variable-speed electric motor 62. Preferably, theangular velocity of the rotor is controlled by a torque converter ormagnetic clutch 63. The primary heat transfer portion of the rotor 68,functioning as the evaporator during the cooling mode, is insulated fromthe secondary heat transfer position functioning as the condenser. A fan64, rotated by a variable speed motor 65, is provided for movingatmospheric air over the secondary heat transfer portion of the rotor. Ablower wheel 66 inside a blower housing rotated by a variable speedmotor 67, is provided for moving air over the primary heat transferportion of the rotor creating air circulation in the primary heatexchange circuit.

[0209] As shown, the air temperature at the inlet of the secondary heatexchanging chamber 14 is sensed by a temperature sensor located in theair flow upstream of the secondary heat transfer portion 69, whereas theair temperature at the outlet thereof is sensed by a temperature sensorlocated in the air flow downstream from the secondary heat transferportion 69. The air temperature at the inlet of the primary heatexchanging chamber 13 is sensed by a temperature sensor located in theair flow upstream of the primary heat transfer portion 68, wherein theair temperature at the outlet thereof is sensed by a temperature sensorlocated downstream from the primary heat transfer portion 68. A simpleexternal on/off thermostat switch 9 can be used to measure temperature Tand thus start motors 62, 65 and 67 during the heating or cooling modeof operation.

[0210] During the cooling mode of operation, the function of the airsupply duct 60 is to convey refrigerated (i.e. cooled/conditioned) airfrom the primary heat transfer portion of the rotor, into the structure(e.g. space to be cooled), whereas the function of the air return duct61 is to convey air from the structure back to the primary heat transferportion for cooling. During the heating mode of operation, the directionof the rotor is reversed by torque generator 62, and the function of theair supply duct is to convey heated air from the primary heat transferportion of the rotor, into the structure (e.g. space to be heated),whereas the function of the air return duct 61 is to convey air from thestructure back to the primary heat transfer portion for heating.

[0211] Second Illustrative Embodiment Of Heat Transfer Engine Hereof

[0212] With reference to FIGS. 14A through 15L, the second illustrativeembodiment of the heat transfer engine of the present invention will bedescribed in detail.

[0213] As shown in FIG. 14A, the heat transfer engine of the secondillustrative embodiment 70 comprises a stator housing 71 within which aturbine-like rotor 72 is rotatably supported. As shown, the rotor isrealized as a solid rotary structure having a turbine-like geometry.Within the rotor structure, a closed self-circulating fluid-carryingcircuit 73 is embodied. As in the first illustrative embodiment, theclosed fluid carrying circuit has spiraled primary and secondary tubularheat transfer passageways, and a metering device which will be describedin greater detail. However, unlike the first illustrative embodiment,these passageways are molded and/or machined in substantially similardisks of different diameters that are stacked and fastened together toform a unity structure. As shown, heat transfer fins are added to eachof the disks in order to (1) increase the secondary heat transfersurface areas thereof and (2) provide a means of systematic fluidcirculation.

[0214] As shown in FIG. 14B, the stator assembly 70 comprises a pair ofsplit-cast housing halves 71A and 71B which are machined to form thefluid flow circuit, and bolted together with bolts 74. As shown, thestator housing has primary and secondary heat exchanging chambers 75 and76, within which the primary and secondary portions of the heatingexchanging rotor are housed. In order that primary and secondary heatexchanging circuits can be appropriately (i.e. thermally) coupled to theprimary and secondary heat exchanging chambers of the stator housing,respectively, flanged fluid piping couplings (i.e. port connections) 77Aand 77B and 78A and 78B are provided to the input and output ports ofthe primary and secondary heat exchanging chambers of the statorhousing, respectively, as shown in FIGS. 14A, 14B and 20. Conventionalfluid carrying pipes with flanged fittings can be easily connected tothese flanged port connections. As shown, when a pressurized heatexchanging fluid (flowing within primary chamber, it will flow overturbine fins 79A on the primary heat exchanging portion of the rotor,impart torque thereto, and thereafter flow out the output port 77B ofthe primary heat exchanging chamber. Similarly, when a pressurized heatexchanging fluid flowing within the secondary heat exchanging circuit isprovided at the input port 78A of the secondary heat exchanging chamber,it will flow over turbine fins 79B on the secondary heat exchangingportion of the rotor, impart torque thereto, and thereafter flow out theoutput port 78B of the secondary heat exchanging chamber.Understandably, the flow of heat exchanging fluid into the input portsof the primary and secondary heat exchanging chambers of the statorhousing will be such that each such fluid flow imparts torque to therotor shaft in a cooperative manner, to perform positive work. As willbe shown hereinafter, the angular velocity of the rotor can becontrolled in a number of different ways depending on the application athand.

[0215] Referring now to FIGS. 15A through 15L, the structure of therotor of the second illustrative embodiment will be described in greaterdetail.

[0216] As shown in FIGS. 15A, 15B, and 15C the primary heat exchangingportion of the rotor comprises a first set of rotor disks 80A havingradially varying outer diameters and a second set of rotor disks 80Bhaving radially uniform outer diameters. Similarly, the secondary heatexchanging portion of the rotor comprises a first set of rotor disks 81Ahaving radially varying outer diameters and a second set of rotor disks81B having radially uniform outer diameters. As shown in FIG. 15B, eachof these rotor disks has a central bore 82 of substantially the samediameter, and a small section of the fluid flow circuit (i.e.passageway) 83 machined, molded or otherwise formed therein. The exactgeometry of each section of fluid flow passageway within each rotor discwill vary from rotor disk to rotor disk. However, these sections offluid flow passageways combine over the length of the rotor to form thegreater portion of the closed fluid flow circuit 83 embodied within therotor structure of the second illustrative embodiment.

[0217] As shown in FIGS. 15A, 15B, and 15C the central bearing structure80 of the rotor comprises an assembly of subcomponents, namely: an outercylindrically-shaped bearing sleeve 81 for rotational support within asuitable support structure provided within the stator housing; an innerfluid flow cylinder 82 of substantially cylindrical geometry adapted tobe received within bearing sleeve 81, having first and seconddisc-receiving collars 83 and 84 of reduced diameter adapted for receiptby inner rotor disc 85 and 86, respectively; a pair of thrust plates 87and 88 having inner central bores with diameters slightly greater thanthe outer diameter of the inner fluid flow cylinder; and a inner fluidflow tube 89 having a inner bore 90 extending along its entire length,and a spirally-extending flange 91 formed on the exterior surfacethereof, for directing return refrigerant. As will be described ingreater detail hereinafter, the central portion of the rotor functionsnot only as a rotor bearing structure, but also as (i) the refrigerantmetering (i.e. throttling) device of the rotor and (ii) a fluid flowreturn passageway. In order to understand how the subcomponents of thecentral portion of the rotor are interconnected and cooperate to carryout the functions of the rotor, it is necessary to first describe thefiner details of this portion of the rotor structure.

[0218] As shown in FIGS. 15B and 15D, the endmost turbine disks 92 and93 have machined within their plate or body portion, a section of fluidflow passageway 82 which extends from a direction substantiallyperpendicular to the rotor axis of rotation, to a directionsubstantially co-parallel with the rotor axis. These sections of closedfluid flow circuit allow refrigerant to flow continuously from thelinear portion thereof to the spiral portions thereof. Also, in orderthat refrigerant can be added or removed from the fluid flow circuit ofthe rotor, each end turbine disk is provided with a charging port 94which is in fluid communication with its central bore 82. As shown, theend of turbine disc 92 and 93 have exterior threads 95 which arereceived by matched interior threads on charging port caps 96A and 96Bwhich can be easily screwed onto and off the charging ports of theserotor discs. To prevent refrigerant leakage, a seal 97 is providedbetween each charging port cap and its end rotor disc, as shown.

[0219] As shown in FIGS. 15B, 15E, 15F, and 15G, each turbine disc set,80A and 81A, carries a plurality of turbine-like fins 99 for the purposeof imparting torque to the rotor when heat exchanging fluid flowsthereover while flowing through the heat exchanging chambers of theengine. In general, the shape of these fins will be determined by theirfunction. For example, in particular embodiments where water flow isused to rotate the rotor within the stator housing, the fins will have3-D surface characteristics which aid in imparting hydrodynamicallygenerated torque to the rotor during engine operation. In order to mountthese fins to the rotor discs, each fin has a base portion 100 which isdesigned to be received within a mated slot 101 formed in the outer endsurface of each rotor disc. Various types of techniques may be employedto securely retain these turbine-like fins within their mounting slots.

[0220] As best shown in FIGS. 15E and 15G, the section of fluid flowpassageway machined in the planar body portion of each rotor disk willvary in geometrical characteristics, depending on the location of therotor disc along the rotor axis. As shown, the fluid flow passageway 83in each rotor disk extends about the center of the rotor disc. Notably,rotor discs 85 and 86 are structurally different than the other discscomprising the heat exchanging portions of the rotor of the secondillustrative embodiment. As shown in FIGS. 15H through 15K, inlet andoutlet rotor discs 85 and 86 are machined so that during the coolingmode, refrigerant in vapor state is transported from the first heatexchanging portion of the rotor to the second heat exchanging portionthereof by way of the spiraled passageway 102, and during the heatingmode, vapor refrigerant is transported in the reverse flow directionthrough the central portion of the rotor. In order to achieve such fluidflow functions, the section of fluid passageway in rotor disks 85 and 86must extend radially inward towards enlarged central recesses 91A and91B respectively, which are adapted to receive the end of cylindricalflanges 83 and 84 of fluid flow cylinder 80 shown in FIG. 15B. Like allother rotor disks, inlet and outlet rotor disks 85 and 86 have centralbores 82 which are aligned with the central bore of the other rotordisks in the rotor structure.

[0221] As best shown in FIGS. 15B and 15C, the inner fluid flow cylinder80 has an axial bore machined, or otherwise drilled and formed, alongits longitudinal extent. Also, fluid flow openings 103 and 104 areformed in the cylindrical flange structures 83 and 84, respectively,extending from the end portions of the inner fluid cylinder. Preferably,the inner diameter of the axial bore 105 formed through outer fluid flowcylinder 82 is about 0.002 inches smaller than the outer diameter of theinner fluid flow tube 89 which carries the spirally extending flange 91.Thus when the inner fluid flow tube 89 is installed within the outerfluid flow cylinder 82, as shown in FIG. 15C, a thin, annular-shapedfluid flow channel 102 is formed therebetween along the entire lengththereof. Thus, when subcomponents of the rotor central portion arecompletely assembled, the following relations are established. First,the fluid flow openings 103 and 104 in the flanges of outer fluid flowcylinder 82 are aligned with the terminal portions of the section of thefluid flow passageway in inlet and outlet rotor discs 85 and 86 (i.e. atthe circumferential edge of circular recess 91A and 91B formed in thesedisc sections). Then the annular-shaped fluid flow channel 102 placesthe portion of the fluid flow circuit along the first heat exchangingportion of the rotor in fluid communication with the portion of thefluid flow circuit along the second heat exchanging portion of therotor. Ultimately, fluid flow continuity is established between the endrotor discs 92 and 93 along the rotor axis by the linear flow passageway82 that is realized by the piecewise assembly of the central boresformed in each rotor disc and the bore 90 formed through inner fluidflow tube 89 in the central portion of the rotor. The above-describedstructural features of the rotor of the second illustrative embodimentensures continuity along the entire fluid flow passageway within theclosed fluid flow circuit embodied within the rotor.

[0222] As will be described in greater detail hereinafter, the sectionof fluid flow passageway 90 passing through the inner fluid flow tube 89functions as a bi-directional throttling (i.e. metering) device withinthe rotor, as it serves to effectively restrict the flow of refrigerantpassing therethrough by virtue of its length and inner diametercharacteristics. Based on the refrigerant used within the rotor andexpected operating pressure and temperature conditions, the length andinner diameter dimensions of the linear flow passageway through theinner fluid flow tube (i.e. throttling channel) can be selected so thatthe required amount of throttling is provided within the closed fluidcircuit during engine operation. For example, assuming it is desired todesign one-quarter horsepower (¼HP) heat transfer engine with a capacityof 11,310 BTUH, and the linear length of the throttling channel is aboutfour (4) inches, then assuming a rotor operating temperature of about50° F. and pressure of about 84 PSIG (pounds per square inch gauge)utilizing monochlorofluoromethane refrigerant (R22), the diameter ofthrottling channel will need to be about 0.028 inches. Depending on thetotal internal volume of the self-circulating fluid flow circuit withinthe rotor, the total refrigerant charge required can be as little as 1.5pounds of liquid refrigerant for small capacity systems, to hundreds ofpounds of liquid refrigerant for larger capacity systems. As the numberof rotor disks is increased, the total internal volume of the closedfluid flow circuit will be increased, and so too the amount ofrefrigerant that must be charged into the system. In principle, therotor structure described above can be made using virtually any numberof rotor disks. It is understood, however, that the number of rotordisks used will depend, in large part, on the thermal load requirements(tonnage in BTUH) which must be satisfied in the application at hand.

[0223]FIG. 15A shows the assembled rotor structure of the secondillustrative embodiment removed from within its stator. This figuresshows the secondary heat transfer portion, primary heat transferportion, the rotor shaft 80, the rotor fins 99, and charging ports 95and 96 of the rotor. The assembly of the rotor structure of the secondillustrative embodiment may be achieved in a variety of ways. Forexample, once assembled in their proper order and configuration, therotor disks can be welded together and thus avoid the need forpressure/liquid-seals (e.g. gaskets), or bolted together and thusrequire the need for seals or gaskets. In alternative embodiments,portions of the rotor structure may be realized using casted parts whichcan be assembled together using welding and/or bolting techniques wellknown in the art.

[0224] Heat Transfer Process of the Second Embodiment

[0225] Referring to FIGS. 16A to 16HF, the refrigeration process of thepresent invention will now be described with the heat transfer engine ofthe second illustrative embodiment in its cooling mode of operation.Notably, each of these drawings schematically depicts, from across-sectional perspective, both the first and second heat exchangingportions of the rotor. This presentation of the internal structure ofthe closed fluid flow passageway throughout the rotor provides a clearillustration of both the location and the state of the refrigerant alongthe closed fluid flow passageway thereof. As will be apparenthereinafter, the heat transfer engine turbine of the second illustrativeembodiment, like the heat transfer engine of the first embodiment,accomplishes a refrigeration affect through the sub-processes ofthrottling, evaporation, superheating, vapor compression,desuperheating, condensation, liquid seal formation and liquidpressurization in the same order except using the turbine-like rotorstructure described above.

[0226] In FIG. 16A, the rotor is shown at its rest position, which isindicated by the absence of any rotational arrow about the rotor shaft.At this stage of operation, the internal volume of the closed fluidcircuit is occupied by about 65% of refrigerant in its liquid state. Theentire spiral return passageway along the rotor shaft is occupied withliquid refrigerant, while the heat exchanging portions of the rotor areoccupied with liquid refrigerant at a level set by gravity in the normalcourse. No throttling of liquid into refrigerant vapor occurs at thisstage of operation. The portion of the fluid passageway above the liquidlevel in the rotor is occupied by refrigerant in a gaseous state. Theclosed fluid flow passageway is thoroughly cleaned and dehydrated priorto the addition of the selected refrigerant to prevent any contaminationthereof.

[0227] As shown in FIG. 16B, the rotor is rotated in a clockwise (CW)direction within the stator housing of the heat transfer engine. At thisstage of operation, the liquid refrigerant within the spiraledpassageway of the shaft begins to flow into the secondary heat transfer(i.e. exchanging) portion of the rotor and occupies substantially theentire volume thereof. At this start-up stage of operation, throttlingof liquid refrigerant into vapor refrigerant begins to occur across thethrottling channel bore 90 inside the rotor. while the rotor continuesto rotate in a clockwise (CW) direction with increasing angularvelocity, the Liquid Seal moves towards the secondary heat transferportion, while refrigerant flowing into the primary heat transferportion of the rotor is restricted by the throttling channel and thusliquid refrigerant accumulates within the secondary heat transferportion thereof. At this stage of operation, very little refrigerantflows into the primary heat transfer portion of the rotor, and thus norefrigeration affect has yet taken place. The small amount ofrefrigerant vapor present in the primary heat transfer portion of therotor will acquire some Superheat which, as a result of the Liquid Seal,will be retained in the vapor and gaseous refrigerant in the primaryheat transfer portion of the rotor.

[0228] As shown in FIG. 16C, the rotor continues to increase in angularvelocity in the CW direction. At this stage of operation, the LiquidPressurization Length of the refrigerant begins to create enoughpressure within the secondary heat transfer portion of the rotor toovercome the pressure restriction presented by the throttling channel,and thus liquid refrigerant begins to flow into the primary heattransfer portion of the rotor. As shown in FIG. 16C, the Liquid Seal hasmoved along the rotor shaft towards the secondary heat transfer portionof the rotor thereof. At this stage of operation, refrigerant beyond thethrottling channel and extending into about the first spiral of fluidflow passageway within the primary heat transfer portion, is in the formof a homogeneous fluid (i.e. a mixture of refrigerant in both its liquidand vapor state). The homogeneous fluid entering the primary heattransfer portion of the rotor “displaces” the gaseous refrigeranttherewithin, thereby pushing it downstream into the spiraled passagewayof the rotor shaft. Sufficient throttling of liquid refrigerant intovapor occurs causing a sufficient temperature drop in the primary heattransfer portion of the rotor and thus causing transfer of Superheatinto the gaseous refrigerant. A “cooler” vapor created by the throttlingprocess of enters the primary heat transfer portion of the rotor andbegins to absorb more Superheat. Refrigerant gas and vapor arecompressed between (i) the homogeneous fluid in the primary heattransfer portion and (ii) the Liquid Seal formed along the spiraledfluid flow passageway of the rotor shaft.

[0229] Notably, at the stage of operation shown in FIG. 16C, there isonly enough pressure in the secondary heat transfer section of the rotorto cause a minimal amount of liquid refrigerant to flow into the primaryheat transfer portion thereof, and thus only slight throttling (i.e.evaporation) of liquid refrigerant into vapor occurs. At this stage,some vapor is beginning to form downstream in the primary heat transferportion of the rotor; however, this is really “flash” gas produced fromthe throttling process. Consequently, at this stage of operation, theonly heat being absorbed by the refrigerant is Superheat in theSuperheat Length of the refrigerant stream, and thus refrigeration hasonly begun to occur. At this stage of the heat transfer process, aLiquid Line is formed in where the homogeneous fluid ends and the vaporbegins along the length of the primary heat transfer portion. Asillustrated in FIGS. 16C through 16E, the Liquid Line can occupy (i.e.manifest itself along) a short length of the primary heat transferportion as a mixture of homogeneous fluid and a very dense vapor whichextends downstream to the Superheat Length. The exact location of theLiquid Line along the primary heat transfer portion of the rotor willvary depending on the quantity of homogeneous fluid therein, which willbe proportional to the amount of heat being absorbed and the thermalload imposed on the primary heat transfer portion of the rotor.

[0230] As the rotor continues to increase its angular velocity in theclockwise (CW) direction towards steady state speed, as shown in FIG.16D, the amount of refrigerant vapor in the primary heat transferportion increases due to increased throttling and production of “Flash”gas as a result of the same. The effect of this vapor increase is anincrease in the quantity of homogeneous fluid entering the primary heattransfer portion of the rotor. At this stage of the process the LiquidSeal has moved even further along the rotor shaft towards the secondaryheat transfer portion. Also, less liquid refrigerant occupies thespiraled passageway of the rotor shaft, while more homogeneous fluidoccupies the primary heat transfer portion of the rotor. As indicated,at this stage of operation, the direction of heat flow (i.e. in the formof Superheat) is from the primary heat transfer portion of the rotor tothe secondary heat transfer portion thereof. However at this stage ofoperation, this heat flow is trapped behind the Liquid Seal formed alongthe spiraled passageway of the rotor shaft.

[0231] As the rotor continues to further increase angular velocity inthe clockwise (CW) direction towards its steady state speed as shown inFIG. 16E, the quantity of refrigerant vapor within the primary heattransfer portion of the rotor continues to increase due to the increasedproduction of flash gas from throttling of liquid refrigerant across thethrottling channel. During this stage of operation, the Liquid Seal hasmoved towards the end of the rotor shaft and the secondary heat transferportion inlet thereof. Also, the flow of heat (i.e. in the form ofSuperheat) from the primary heat transfer portion is still trappedbehind the Liquid Seal in the spiraled passageway of the rotor shaft.Consequently, the Superheat from the primary heat transfer portion ofthe rotor is unable to pass onto the secondary heat transfer portion ofthe rotor. Consequently, optimal operation is not yet achieved at thisstage of engine operation. During this stage of operation some heat(Superheat) may transfer into the rotor shaft from the refrigerant vaporif the shaft temperature is less that the temperature of the refrigerantvapor; and some heat may transfer into the refrigerant vapor if therefrigerant vapor temperature is less than that of the rotor shaft.

[0232] The rotor shaft and its internal spiraled passageway provideprimary and secondary Superheat transfer surfaces where heat can beeither absorbed into or discharged from the vapor stream circulatingwithin the closed fluid flow circuit of the rotor. Heat produced byfriction from the rotor shaft bearings is absorbed by the refrigerantvapor along the length of the rotor shaft and can add to the amount ofSuperheat entering the secondary heat transfer portion. This additionalSuperheat further increases the temperature difference between theSuperheated vapor and the secondary heat transfer surfaces of thesecondary heat transfer portion. In turn, this increases the rate ofheat flow from the Superheated vapor within the rotor, and thus enhancesthe heat transfer locations required to achieve steady state operation.

[0233] At the stage of operation shown in FIG. 16E, the rotor isapproaching, but has not yet attained its steady-state angular velocity,which as shown in performance characteristics of FIG. is referred to as“Minimal Velocity” or “Threshold Velocity”. Consequently, the heattransfer engine is not yet operating along the linear portion of itsoperating characteristic. As shown in FIG. 16F, the remaining liquidrefrigerant in the rotor shaft is now completely displaced byrefrigerant vapor produced as a result of the evaporation of the liquidrefrigerant in primary heat transfer portion of the rotor. Consequently,Superheat produced from the primary heat transfer portion of the rotoris permitted to flow through the spiraled passageway of the rotor shaftand into the secondary heat transfer portion, where it can be liberatedby way of condensation across the secondary heat transfer portion. Asshown, Superheat Length of the refrigerant stream within primary heattransfer portion of the rotor has decreased in effective length, whilethe Evaporation Length of the refrigerant stream has increasedproportionally, indicating that the refrigeration effect within theprimary heat transfer portion is increasing towards the Balanced Pointor steady state condition. At this stage of operation, the Liquid Sealis no longer located along the rotor shaft, but within the secondaryheat transfer portion of the rotor, near the end of the rotor shaft.Vapor compression has begun to occur in the tail end of the primary heattransfer portion and along the spiraled passageway of the rotor. At thisstage of operation, the pressure of the liquid refrigerant along theLiquid Pressurization Length has increased sufficiently enough tofurther increase the production of homogeneous fluid in the primary heattransfer portion of the rotor. This also causes the quantity of liquidin the secondary heat transfer portion to decrease the “Pulling Effect”on the flash gas and vapor located in the spiraled passageway in therotor shaft, as well as in the primary heat transfer portion of therotor downstream from the homogeneous fluid. The pulling affect on theflash gas enhances vapor compression taking place along the spiraledpassageway of the rotor shaft. At this stage of operation thehomogeneous fluid is evaporating absorbing heat within the primary heattransfer portion of the rotor for transference and systematic dischargefrom the secondary heat transfer portion. In other words, during thisstage of operation, the vapor within the primary heat transfer portionof the rotor can contain more Superheat by volume than the gas withwhich it is mixed. Thus, the increased volume in dense vapor in theprimary heat transfer portion provides a means of storing Superheat(absorbed from the primary heat exchanging circuit) until the vaporstream flows into the secondary heat transfer portion of the rotor whereit can be liberated to the secondary heat exchanging circuit by way ofconduction.

[0234] As shown in FIG. 16F, the heat transfer engine of the presentinvention is shown operating at what shall be called the “Balance PointCondition” (i.e. steady-state condition). At this stage of operation,the refrigerant within the rotor has attained the necessary phasedistribution where simultaneously there is an equal amount ofrefrigerant being evaporated in the primary heat transfer portion asthere is refrigerant vapor being condensed in the secondary heattransfer portion of the rotor. At this stage of operation, the heattransfer engine is operating along the linear portion of its operatingcharacteristic, shown in FIG. 9. At this stage, there exists a range orband of angular velocities within which the rotor can rotate and a rangeof loading conditions within which the rotor can transfer heat whilemaintaining a substantially linear relationship between (i) the rate ofheat transfer between the primary and secondary heat exchanging portionsof the rotor and the (ii) angular velocity thereof. Outside of thisrange of operation, these parameters no longer follow a linearrelationship. This has two major consequences. The first consequence isthat the control structure (i.e. system controller) of the engineperforms less than ideally. The second consequence is that maximalrefrigeration cannot be achieved.

[0235] As shown in FIG. 16G, the Superheat that has “accumulated” in therefrigerant vapor during the start up sequence shown in FIGS. 16Athrough 16F begins to dissipate from the DeSuperheat Length of therefrigerant stream along the secondary heat transfer portion of therotor. At this stage of operation, the density of the refrigerant gasincreases while vapor compression occurs as a result of Superheat beingcarried by the refrigerant gas from the Superheat Length along theprimary heat transfer portion to the DeSuperheat Length along thesecondary heat transfer portion via the spiraled passageway of the rotorshaft. Thus, as the Superheat is dissipated in the secondary heattransfer portion of the rotor and compressed vapor in the secondary heattransfer portion thereof begins to condense into liquid refrigerant, adenser 1 vapor remains. Consequently, the spiraled passageway of therotor shaft has a greater compressive affect on the vapor therein atthis stage of operation. In other words, the spiraled passageway of theshaft pressurizes the superheated gas and dense vapor against the LiquidSeal formed in the secondary heat transfer portion of the rotor.

[0236] As shown in FIG. 16G, pressurization of liquid refrigerant in thesecondary heat transfer portion of the rotor pushes the liquidrefrigerant through the throttling device at a sufficiently higherpressure, which causes a portion of the liquid refrigerant to “flash”into a gas. This reduces the temperature of the remaining homogeneousfluid (liquid and dense vapor) entering the primary heat transferportion thereof. The liquid refrigerant portion of the homogeneousfluid, in turn, evaporates creating sufficient vapor pressure thereinwhich displaces vapor downstream within the primary heat transferportion, into the spiraled passageway of the rotor shaft. This vaporpressure, enhanced by vapor compression caused by the spiraledpassageway in the rotor shaft, pushes the produced vapor into thesecondary heat transfer portion of the rotor, where its Superheat isliberated over the DeSuperheat Length of the refrigerant stream.

[0237] At the Balance Point condition, a number of conditions remainthroughout steady-state operation. Foremost, the Liquid Seal tends toremain near the same location in the secondary heat transfer portion ofthe rotor, while the Liquid Line tends to remain near the same locationin the primary heat transfer portion thereof. Secondly, the temperatureand pressure of the (refrigerant in the secondary heat transfer portionof the rotor is higher than the refrigerant in the primary heat transferportion thereof. Thirdly, the rate of heat transfer from the primaryheat exchanging chamber of the engine into the primary heat transferportion thereof is substantially equal to the rate of heat transfer fromthe secondary heat transfer portion of the engine into the secondaryheat exchanging chamber thereof. Thus, if the primary heat transferportion of the rotor is absorbing heat at about 12,000 BTUH, then thesecondary heat transfer portion thereof is dissipating about 12,000BTUH.

[0238] Applications Of Second Embodiment Of Heat Transfer Engine Hereof

[0239] In FIG. 17, a heat transfer system according to the presentinvention is shown, wherein the rotor of the heat transfer enginethereof 70 is driven (i.e. torqued) by fluid flow streams 95A flowingthrough the secondary heat exchanging circuit 95B of the system. In thisheat transfer system, heat liberated from the secondary heat exchangingportion 94 of the rotor is absorbed by a fluid 95A from pump 97A and atypical condenser cooling tower 97. As shown, cooling tower 97 is partof systematic fluid flow circuit in a cooling tower piping system whereheat is exchanged with the cooling tower and consequently with theambient atmosphere. As shown in FIG. 17, the heat transfer engine 70 is“pumping” a fluid 96A, such as water, through a typical closed-loop tubeand shell heat exchanger 98 and its associated piping 96B and flowcontrol valve 98A. This heat transfer system is ideal for use inchilled-water air conditioning systems as well as process-water coolingsystems.

[0240] As shown in FIG. 17, the fluid flow rate controller in primaryheat exchanging circuit 96B is realized as a flow control valve 98Awhich receives primary heat exchanging fluid 96A by way of the primaryheat exchanging portion 93 of the heat exchanging engine 70. The systemcontroller 11 generates suitable signals to control the operation of theflow control valves (i.e. by adjusting the valve flow aperture diameterduring engine operation). Preferably, in the secondary heat exchangingcircuit 95B, the secondary fluid flow rate controller is realized as aflow rate control valve 97B designed for controlled operation under thecontrol of system controller 11.

[0241] In FIG. 18, a modified embodiment of heat transfer system of FIG.17 is shown. The primary difference between these systems is that thefluid inlet and outlet ports 77A and 77B of the system shown in FIG. 18are arranged on the same side of the engine, and the rotor shaft 77thereof is extended beyond the stator housing to permit an externalmotor 98 to drive the same in either direction of rotation using atorque converter 99.

[0242] In FIG. 19, another embodiment of a heat transfer systemaccording to the present invention is shown, wherein two (or more)turbine-like heat transfer engines 125 and 127 are connected in acascaded manner. As shown, the primary heat transfer portion of heattransfer engine 125 is in thermal communication with the secondary heattransfer portion of heat transfer portion 127, while the primary heattransfer portion of the rotor of engine 127 is in thermal communicationwith a closed chilled water loop flowing through the primary heatexchanging chamber thereof, and the secondary heat transfer portion ofthe rotor of engine 125 is in thermal communication with a closedprocess-water loop flowing through the secondary heat exchanging chamberthereof. As shown, the rotor of heat transfer engine 125 is driven byelectric motor 126 coupled there by way of a first torque converter,while the rotor of heat transfer engine 127 is driven by electric motor128 coupled therebetween by way of a second torque converter.

[0243] In FIG. 20, an alternative embodiment of a heat transfer systemof the present invention is shown, wherein a hybrid-type heat transferengine is employed. As shown, the hybrid-type heat transfer engine has asecondary heat transfer portion 129 adapted from the heat transferengine of the first embodiment and a secondary heat transfer portion 130adapted from the heat transfer engine of the second embodiment. Thefunction of the primary heat transfer portion is to serve as an aircooled condenser, whereas the function of the secondary heat transferportion is to serve as an evaporator in a closed-loop fluid chiller. Asshown in FIG. 20, rotational torque is imparted to the rotor of thehybrid engine by allowing fluid to flow over the primary heat transfervanes of the primary heat transfer portion 130 thereof.

[0244] In FIG. 21, another embodiment of a heat transfer system of thepresent invention is shown, wherein another hybrid-type heat transferengine is employed. As shown, the hybrid-type heat transfer engine has asecondary heat transfer portion 129 adapted from the heat transferengine of the first embodiment and a secondary heat transfer portion 130adapted from the heat transfer engine of the second embodiment. Thefunction of the primary heat transfer portion is to serve as an airconditioning evaporator, whereas the function of the secondary heattransfer portion is to serve as a condenser in an open loop fluid cooledcondenser. As shown in FIG. 21, rotational torque is imparted to therotor of the hybrid engine by an electric motor 134 connector to therotor shaft 135 by a magnetic torque converter 133, whereas allowingfluid to flow over the primary heat transfer vanes of the primary heattransfer portion 130 thereof.

[0245] Applications Of Either Embodiment Of The Heat Transfer EngineHereof

[0246] In FIG. 22, a heat transfer engine of the present invention isembodied within an automobile. In this application, the rotor of theheat transfer engine is rotated by an electric motor driven byelectrical power which is supplied through a power control circuit, andproduced by the automobile battery that is recharged by an alternatorwithin the engine compartment of the automobile.

[0247] In FIG. 23, a heat transfer engine of the present invention isembodied within a refrigerated tractor trailer truck. In thisapplication, the rotor of the heat transfer engine is rotated by anelectric motor driven by electrical power which is supplied through apower control circuit and produced by a bank of batteries recharged byan alternator within the engine compartment of the truck.

[0248] In FIG. 24, a plurality of heat transfer engines of the presentinvention are embodied within an aircraft. In this application, therotor of each heat transfer engine is rotated by an electric motor. Theelectric motor is driven by electrical power which is produced by anonboard electric generator and supplied to the electric motors throughvoltage regulator and temperature control circuit.

[0249] In FIG. 25, a plurality of heat transfer engines of the presentinvention are embodied within a refrigerated freight train. In thisapplication, the rotor of each heat transfer engine is rotated by anelectric motor driven by electrical power. The electric power isproduced by an onboard pneumatically driven electric generator, and issupplied to the electric motors through a voltage regulator andtemperature control circuit.

[0250] In FIG. 26, a plurality of heat transfer engines of the presentinvention are embodied within a refrigerated shipping vessel. In thisapplication, the rotor of each heat transfer engine is rotated by anelectric motor driven by electrical power. The electric power isproduced by an onboard pneumatically driven electric generator, and issupplied to the electric motors through a voltage regulator andtemperature control circuit.

[0251] Having described various illustrative embodiments of the presentinvention, various modifications readily come to mind.

[0252] Various embodiments of the heat transfer engine technology of thepresent invention have been described above in great detail above.Preferably, each embodiment is designed using 3-D computer workstationhaving 3-D geometrical modeling capabilities, as well as mathematicalmodeling tools to develop mathematical models of each engine hereofusing equation of energy, equations of motion and the like, well knownin the fluid dynamics and thermodynamics art. Using suchcomputational-based models, simulation of proposed system designs can becarried out on the computer workstation, performance criteriaestablished, and design parameters modified to achieve optimal heattransfer engine designs based on the principles of the present inventiondisclosed herein.

[0253] The illustrative embodiments described in detail herein havegenerally focused on cooling or heating fluid (e.g. air) flow streamspassing through the primary heat exchanging circuit to which the heattransfer engines hereof are operably connected. However, in someapplications, such as dehumidification, it is necessary to both cool andheat air using one or more heat transfer engines of the presentinvention. In such applications, the air flow (being conditioned) can beeasily directed over the primary heat exchanging portion of the rotor inorder to condense moisture in the air stream, and thereafter directedover the secondary heat exchange portion of the rotor in order tore-heat the air for redistribution (reentry) into the conditioned spaceassociated with the primary heat exchanging fluid circuit. Using suchtechniques, the heat transfer engines described hereinabove can bereadily modified to provide engines capable of performing both coolingand heating functions.

[0254] In general, both the coiled heat transfer engine and theembedded-coil (i.e. turbine line) heat transfer engine turbine of thepresent invention can be cascaded in various ways, utilizing variousrefrigerants and fluids, for various capacity and operating temperaturerequirements. Digital or analog type temperature and pressure sensorsmay be used to realize the system controllers of such embodiments. Also,electrical, pneumatic, and/or hydraulic control structures (or anycombination thereof) can also be used to realize such embodiments of thepresent invention.

[0255] Although preferred embodiments of the invention have beendescribed in the foregoing Detailed Description and illustrated in theaccompanying drawings, it will be understood that the invention is notlimited to the embodiments disclosed, but is capable of numerousrearrangements, modifications, and substitutions of parts and elementswithout departing from the spirit of the invention. Accordingly, thepresent invention is intended to encompass such rearrangements,modifications, and substitutions of parts and elements as fall withinthe scope and spirit of the accompanying Claims to Invention.

What is claimed is:
 1. A heat transfer engine for transferring heatbetween first and second heat exchanging circuits, comprising: astationary housing having first and second heat transfer chambers, and athermal isolation barrier disposed therebetween, said first and secondheat transfer chambers each having first and second ports and acontinuous passageway therebetween; and a rotatable heat transferstructure rotatably supported within said stationary housing about anaxis of rotation and having a substantially symmetrical moment ofinertia about said axis of rotation, said rotatable heat transferstructure having a first end portion disposed within said first heattransfer chamber, a second end portion disposed within said second heattransfer chamber, and an intermediate portion disposed between saidfirst and second end portions and including a means for, said rotatableheat transfer structure embodying a closed fluid circuit symmetricallyarranged about said axis of rotation, and having a return portionextending along the direction of said axis of rotation and at least asubportion of said return portion having a helical geometry, and aninterior volume for containing a predetermined amount of a heat carryingmedium contained within said closed fluid circuit which automaticallycirculates within said closed fluid circuit as said rotatable heattransfer structure is rotated about said axis of rotation and therewhileundergoes a phase transformation within said closed fluid circuit inorder to carry out a heat transfer process between said first and secondportions of said rotatable heat transfer structure, said first endportion of said rotatable heat transfer structure being disposed inthermal communication with said first heat exchanging circuit, saidsecond end portion rotatable heat transfer structure being disposed inthermal communication with said second heat exchanging circuit, saidintermediate portion being physically adjacent to said thermal barrierso as to present a substantially high thermal resistance to heattransfer between said first and second heat transfer chambers duringoperation of said heat transfer engine, and said heat carrying mediumbeing characterized by a predetermined heat of evaporation at which saidheat carrying medium transforms from liquid phase to vapor phase, and apredetermined heat of condensation at which said heat carrying mediumtransforms from vapor phase to liquid phase, and wherein the directionof phase change of said heat carrying liquid is reversible; and a flowrestriction means disposed along said intermediate portion forrestricting the flow of said heat carrying fluid through said closedfluid circuit as said rotatable heat transfer structure is rotatedwithin about said axis of rotation;.
 2. The heat transfer engine ofclaim 1, which further comprises: torque generation means for impartingtorque to said rotatable heat transfer structure and causing saidrotatable heat transfer structure to rotate about said axis of rotation;and torque control means for controlling said torque generating means inresponse to the temperature of said heat exchanging medium sensed atsaid inlet and outlet ports in said first and second heat transferchambers.
 3. The heat transfer engine of claim 2, wherein said torquegenerating means comprises: a motor having a drive shaft operablyconnected to said rotatable heat transfer structure, wherein the angularvelocity of said drive shaft is maintained within a predetermined rangeof angular velocity by said torque controlling means.
 4. The heattransfer engine of claim 2, wherein said torque generating meanscomprises turbine blades disposed on at least one of said first andsecond end portions of said rotatable heat transfer structure, such thatsaid turbine blades are imparted torque by a first or second heatexchanging medium flowing through said first or second heat transferchambers during the operation of said heat transfer engine.
 5. The heattransfer engine of claim 2, wherein said torque generating meanscomprises: a steam turbine having a drive shaft operably connected tosaid rotatable heat transfer structure, for imparting torque to saidrotatable heat transfer structure, and wherein said torque controllingmeans comprises means for controlling the angular velocity of the driveshaft of said steam turbine.
 6. The heat transfer engine of claim 1,wherein the first end portion of said rotatable heat transfer structurefunctions as an evaporator and the second end portion of said rotatableheat transfer structure functions as a condenser when said rotatableheat transfer structure rotates in a clockwise direction.
 7. The heattransfer engine of claim 1, wherein the first end portion of saidrotatable heat transfer structure functions as an condenser and thesecond end portion of said rotatable heat transfer structure functionsas an evaporator when said rotatable heat transfer structure rotates ina counter-clockwise direction.
 8. The heat transfer engine of claim 1,wherein said rotatable heat transfer structure comprises a rotor portionhaving a substantially symmetrical moment of inertia about said axis ofrotation, and said closed fluid circuit is realized as athree-dimensional flow passageway of closed loop design formed in saidrotor portion, said three-dimensional flow passageway comprising afirst, second, third and fourth spiral flow passageway portionsconnected in a series configuration about said axis of rotation, in thenamed order.
 9. The heat transfer engine of claim 1, wherein said rotorportion comprises a plurality of rotor discs assembled together to forma unitary structure, wherein each said rotor disc has formed therein asection of grooving which relates to a portion of said three-dimensionalflow passageway formed in said rotor portion.
 10. The heat transferengine of claim 1, wherein said rotatable heat transfer structurecomprises a rotor shaft along which said return portion of said closedfluid circuit extends, and wherein said closed fluid circuit is realizedas a three-dimensional tubing configuration supported about said rotorshaft having a first, second, third and fourth spiral tubing sectionscontinuously connected in a series configuration about said axis ofrotation, in the named order.
 11. The heat transfer engine of claim 1,wherein said return portion has a helical geometry which extendssubstantially along the entire extend of said rotor shaft.
 12. The heattransfer engine of claim 1, which further comprises: first connectionmeans for interconnecting a first heat exchanging circuit to said firstand second ports of said first heat transfer chamber, so as to permit afirst heat exchanging medium to flow through said first heat exchangingcircuit and said first chamber during the operation of said reversibleheat transfer engine; and second connection means for interconnecting asecond heat exchanging circuit to said first and second ports of saidsecond heat transfer chamber, so as to permit a second heat exchangingmedium to flow through said second heat exchanging circuit and saidsecond heat transfer chamber during the operation of said reversibleheat transfer engine, while said first and second heat exchangingcircuits are in substantial thermal isolation of each other.
 13. Theheat transfer engine of claim 12, which further comprises temperaturesensing means for measuring the temperature of said heat exchangingmedium flowing through said inlet and outlet ports of said first andsecondary heat transfer chambers.
 14. The heat transfer engine of claim12, wherein said first heat exchanging medium flow through said firstheat exchanging circuit is air, and said second heat exchanging mediumflow through said second heat exchanging circuit is air.
 15. The heattransfer engine of claim 12, wherein said first heat exchanging mediumflow through said first heat exchanging circuit is water, and saidsecond heat exchanging medium flow through said second heat exchangingcircuit is air.
 16. The heat transfer engine of claim 12, wherein saidfirst heat exchanging medium flow through said first heat exchangingcircuit is water, and said second heat exchanging medium flow throughsaid second heat exchanging circuit is water.
 17. The heat transferengine of claim 12, wherein said first heat exchanging medium flowthrough said first heat exchanging circuit is air, and said second heatexchanging medium flow through said second heat exchanging circuit iswater.
 18. A heat transfer engine for transferring heat between firstand second heat exchanging circuits, comprising: a stationary housinghaving first and second heat transfer chambers, and a thermal isolationbarrier disposed therebetween, said first and second heat transferchambers each having first and second ports and a continuous passagewaytherebetween; and a rotatable heat transfer structure rotatablysupported within said stationary housing about an axis of rotation andhaving a substantially symmetrical moment of inertia about said axis ofrotation, said rotatable heat transfer structure having a first endportion disposed within said first heat transfer chamber, a second endportion disposed within said second heat transfer chamber, and anintermediate portion disposed between said first and second endportions, said rotatable heat transfer structure embodying a closedfluid circuit symmetrically arranged about said axis of rotation, andhaving a return portion extending along the direction of said axis ofrotation, and an interior volume for containing a predetermined amountof a heat carrying medium contained within said closed fluid circuitwhich automatically circulates within said closed fluid circuit as saidrotatable heat transfer structure is rotated about said axis of rotationand therewhile undergoes a phase transformation within said closed fluidcircuit in order to carry out a heat transfer process between said firstand second portions of said rotatable heat transfer structure, saidfirst end portion of said rotatable heat transfer structure beingdisposed in thermal communication with said first heat exchangingcircuit, said second end portion rotatable heat transfer structure beingdisposed in thermal communication with said second heat exchangingcircuit, said intermediate portion being physically adjacent to saidthermal barrier so as to present a substantially high thermal resistanceto heat transfer between said first and second heat transfer chambersduring operation of said heat transfer engine, said heat carrying mediumbeing characterized by a predetermined heat of evaporation at which saidheat carrying medium transforms from liquid phase to vapor phase, and apredetermined heat of condensation at which said heat carrying mediumtransforms from vapor phase to liquid phase, and wherein the directionof phase change of said heat carrying liquid is reversible, and saidrotatable heat transfer structure having predetermined range of angularvelocity over which said heat transfer engine is capable of transferringheat between said first and second end portions of said rotatable heattransferring structure; a flow restriction means disposed along saidintermediate portion for restricting the flow of said heat carryingfluid through said closed fluid circuit; torque generation means forimparting torque to said rotatable heat transfer structure and causingsaid rotatable heat transfer structure to rotate about said axis ofrotation; and torque control means for controlling said torquegenerating means in response to the temperature of said heat exchangingmedium sensed at either said inlet and outlet ports in said first andsecond heat transfer chambers, so that the angular velocity of saidrotatable heat transfer structure is maintained with said predeterminedrange.
 19. The heat transfer engine of claim 18, wherein said torquegenerating means comprises: a motor having a drive shaft operablyconnected to said rotatable heat transfer structure, wherein the angularvelocity of said drive shaft is maintained within a predetermined rangeof angular velocity by said torque controlling means.
 20. The heattransfer engine of claim 18, wherein said torque generating meanscomprises: a motor having a drive shaft operably connected to saidrotatable heat transfer structure, wherein the angular velocity of saiddrive shaft is maintained within a predetermined range of angularvelocity by said torque controlling means.
 21. The heat transfer engineof claim 18, wherein said torque generating means comprises turbineblades disposed on at least one of said first and second end portions ofsaid rotatable heat transfer structure, such that said turbine bladesare imparted torque by a first or second heat exchanging medium flowingthrough said first or second heat transfer chambers during the operationof said heat transfer engine.
 22. The heat transfer engine of claim 18,wherein said torque generating means comprises: a steam turbine having adrive shaft operably connected to said rotatable heat transferstructure, for imparting torque to said rotatable heat transferstructure, and wherein said torque controlling means comprises means forcontrolling the angular velocity of the drive shaft of said steamturbine.
 23. The heat transfer engine of claim 18, wherein the first endportion of said rotatable heat transfer structure functions as anevaporator and the second end portion of said rotatable heat transferstructure functions as a condenser when said rotatable heat transferstructure rotates in a clockwise direction.
 24. The heat transfer engineof claim 18, wherein the first end portion of said rotatable heattransfer structure functions as an condenser and the second end portionof said rotatable heat transfer structure functions as an evaporatorwhen said rotatable heat transfer structure rotates in acounter-clockwise direction.
 25. The heat transfer engine of claim 18,wherein said rotatable heat transfer structure comprises a rotor portionhaving a substantially symmetrical moment of inertia about said axis ofrotation, and said closed fluid circuit is realized as athree-dimensional flow passageway of closed loop design formed in saidrotor portion, said three-dimensional flow passageway comprising afirst, second, third and fourth spiral flow passageway portionsconnected in a series configuration about said axis of rotation, in thenamed order.
 26. The heat transfer engine of claim 18, wherein saidrotor portion comprises a plurality of rotor discs assembled together toform a unitary structure, wherein each said rotor disc has formedtherein a section of grooving which relates to a portion of saidthree-dimensional flow passageway formed in said rotor portion.
 27. Theheat transfer engine of claim 18, wherein said rotatable heat transferstructure comprises a rotor shaft along which said return portion ofsaid closed fluid circuit extends, and wherein said closed fluid circuitis realized as a three-dimensional tubing configuration supported aboutsaid rotor shaft having a first, second, third and fourth spiral tubingsections continuously connected in a series configuration about saidaxis of rotation, in the named order.
 28. The heat transfer engine ofclaim 18, wherein at least a subportion of said return portion has ahelical geometry.
 29. The heat transfer engine of claim 28, wherein saidreturn portion has a helical geometry which extends substantially alongthe entire extend of said rotor shaft.
 30. The heat transfer engine ofclaim 18, wherein said first heat exchanging medium flow through saidfirst heat exchanging circuit is air, and said second heat exchangingmedium flow through said second heat exchanging circuit is air.
 31. Theheat transfer engine of claim 18, wherein said first heat exchangingmedium flow through said first heat exchanging circuit is water, andsaid second heat exchanging medium flow through said second heatexchanging circuit is air.
 32. The heat transfer engine of claim 18,wherein said first heat exchanging medium flow through said first heatexchanging circuit is water, and said second heat exchanging medium flowthrough said second heat exchanging circuit is water.
 33. The heattransfer engine of claim 18, wherein said first heat exchanging mediumflow through said first heat exchanging circuit is air, and said secondheat exchanging medium flow through said second heat exchanging circuitis water.
 34. The heat transfer portion of claim 18, which furthercomprises: first connection means for interconnecting a first heatexchanging circuit to said first and second ports of said first heattransfer chamber, so as to permit a first heat exchanging medium to flowthrough said first heat exchanging circuit and said first chamber duringthe operation of said reversible heat transfer engine; and secondconnection means for interconnecting a second heat exchanging circuit tosaid first and second ports of said second heat transfer chamber, so asto permit a second heat exchanging medium to flow through said secondheat exchanging circuit and said second heat transfer chamber during theoperation of said reversible heat transfer engine, while said first andsecond heat exchanging circuits are in substantial thermal isolation ofeach other.
 35. The heat transfer engine of claim 34, which furthercomprises temperature sensing means for measuring the temperature ofsaid heat exchanging medium flowing through said inlet and outlet portsof said first and secondary heat transfer chambers.
 36. A heat transferengine for transferring heat between first and second heat exchangingcircuits, comprising: a stationary housing having first and second heattransfer chambers, and a thermal isolation barrier disposedtherebetween, said first and second heat transfer chambers each havingfirst and second ports and a continuous passageway therebetween; and arotatable heat transfer structure rotatably supported within saidstationary housing about an axis of rotation and having a substantiallysymmetrical moment of inertia about said axis of rotation, saidrotatable heat transfer structure having a first end portion disposedwithin said first heat transfer chamber, a second end portion disposedwithin said second heat transfer chamber, and an intermediate portiondisposed between said first and second end portions, said rotatable heattransfer structure embodying a closed fluid circuit symmetricallyarranged about said axis of rotation, and having a return portionextending along the direction of said axis of rotation, and an interiorvolume for containing a predetermined amount of a heat carrying mediumcontained within said closed fluid circuit which automaticallycirculates within said closed fluid circuit as said rotatable heattransfer structure is rotated about said axis of rotation and therewhileundergoes a phase transformation within said closed fluid circuit inorder to carry out a heat transfer process between said first and secondportions of said rotatable heat transfer structure, said first endportion of said rotatable heat transfer structure being disposed inthermal communication with said first heat exchanging circuit, saidsecond end portion rotatable heat transfer structure being disposed inthermal communication with said second heat exchanging circuit, saidintermediate portion being physically adjacent to said thermal barrierso as to present a substantially high thermal resistance to heattransfer between said first and second heat transfer chambers duringoperation of said heat transfer engine, and said heat carrying mediumbeing characterized by a predetermined heat of evaporation at which saidheat carrying medium transforms from liquid phase to vapor phase, and apredetermined heat of condensation at which said heat carrying mediumtransforms from vapor phase to liquid phase, and wherein the directionof phase change of said heat carrying liquid is reversible; a flowrestriction means disposed along said intermediate portion forrestricting the flow of said heat carrying fluid through said closedfluid circuit; first connection means for interconnecting a first heatexchanging circuit to said first and second ports of said first heattransfer chamber, so as to permit a first heat exchanging medium to flowthrough said first heat exchanging circuit and said first chamber duringthe operation of said reversible heat transfer engine; second connectionmeans for interconnecting a second heat exchanging circuit to said firstand second ports of said second heat transfer chamber, so as to permit asecond heat exchanging medium to flow through said second heatexchanging circuit and said second heat transfer chamber during theoperation of said reversible heat transfer engine, while said first andsecond heat exchanging circuits are in substantial thermal isolation ofeach other; temperature sensing means for measuring the temperature ofsaid heat exchanging medium flowing through said inlet and outlet portsof said first and secondary heat transfer chambers; torque generationmeans for imparting torque to said rotatable heat transfer structure andcausing said rotatable heat transfer structure to rotate about said axisof rotation; and torque control means for controlling said torquegenerating means in response to the temperature of said heat exchangingmedium sensed at said inlet and outlet ports in said first and secondheat transfer.
 37. The heat transfer engine of claim 36, wherein saidtorque generating means comprises; a motor having a drive shaft operablyconnected to said rotatable heat transfer structure, wherein the angularvelocity of said drive shaft is maintained within said predeterminedrange by said torque controlling means.
 38. The heat transfer engine ofclaim 36, wherein said torque generating means comprises turbine bladesdisposed on at least one of said first and second end portions of saidrotatable heat transfer structure, such that said turbine blades areimparted torque by said first or second heat exchanging medium flowingthrough said first or second heat exchanging circuit and said first orsecond heat transfer chamber during the operation of said heat transferengine.
 39. The heat transfer engine of claim 36, wherein said torquegenerating means comprises; a steam turbine having a drive shaftoperably connected to said rotatable heat transfer structure, forimparting torque to said rotatable heat transfer structure, and whereinsaid torque controlling means comprises means for controlling theangular velocity of the drive shaft of said steam turbine.
 40. The heattransfer engine of claim 36, wherein the first end portion of saidrotatable heat transfer structure functions as an evaporator and thesecond end portion of said rotatable heat transfer structure functionsas a condenser when said rotatable heat transfer structure rotates in aclockwise direction.
 41. The heat transfer engine of claim 36, whereinthe first end portion of said rotatable heat transfer structurefunctions as an condenser and the second end portion of said rotatableheat transfer structure functions as an evaporator when said rotatableheat transfer structure rotates in a counter-clockwise direction. 42.The heat transfer engine of claim 36, wherein the return portion of saidclosed fluid circuit has a helical geometry extending from said firstend portion to said second end portion.
 43. The heat transfer engine ofclaim 36, wherein said rotatable heat transfer structure comprises arotor portion having a substantially symmetrical moment of inertia aboutsaid axis of rotation, and said closed fluid circuit is realized as athree-dimensional flow passageway of closed loop design formed in saidrotor portion, said three-dimensional flow passageway comprising afirst, second, third and fourth spiral flow passageway portionsconnected in a series configuration about said axis of rotation, in thenamed order.
 44. The heat transfer engine of claim 18, wherein at leasta subportion of said return portion has a helical geometry.
 45. The heattransfer engine of claim 44, wherein said return portion has a helicalgeometry which extends substantially along the entire extend of saidrotor shaft.
 46. The heat transfer engine of claim 36, wherein saidfirst heat exchanging medium flow through said first heat exchangingcircuit is air, and said second heat exchanging medium flow through saidsecond heat exchanging circuit is air.
 47. The heat transfer engine ofclaim 36, wherein said first heat exchanging medium flow through saidfirst heat exchanging circuit is water, and said second heat exchangingmedium flow through said second heat exchanging circuit is air.
 48. Theheat transfer engine of claim 36, wherein said first heat exchangingmedium flow through said first heat exchanging circuit is water, andsaid second heat exchanging medium flow through said second heatexchanging circuit is water.
 49. The heat transfer engine of claim 18,wherein said first heat exchanging medium flow through said first heatexchanging circuit is air, and said second heat exchanging medium flowthrough said second heat exchanging circuit is water.
 50. A vehicle withon-board heat transfer capabilities comprising: a platform fortransporting objects; and the heat transfer engine of claim 1 mountedaboard said platform.
 51. The vehicle of claim 51, wherein said platformis either an ground transportable structure, an air supportablestructure, and/or water transportable structure.
 52. A vehicle withon-board heat transfer capabilities comprising: a platform fortransporting objects; and the heat transfer engine of claim 18 mountedaboard said platform.
 53. The vehicle of claim 52, wherein said platformis either an ground transportable structure, an air supportablestructure, and/or water transportable structure.
 54. A vehicle withon-board heat transfer capabilities comprising: a platform fortransporting objects; and the heat transfer engine of claim 36 mountedaboard said platform.
 55. The vehicle of claim 54, wherein said platformis either an ground transportable structure, an air supportablestructure, and/or water transportable structure.
 56. A heat transferengine comprising: a stationary housing having first and second heattransfer chambers; a heat transfer structure rotatably supported withinsaid stationary housing about an axis of rotation; torque generationmeans for imparting torque to said heat transfer structure and causingsaid heat transfer structure to rotate about said axis of rotation; andtorque control means for controlling said torque generating means withina closed control loop during the transfer of heat between said a firstand second heat transfer chambers;.
 57. A method transferring heatbetween first and second heat exchanging circuits, comprising the steps:(a) installing between first and second heat exchanging circuits a heattransfer engine which includes a stationary housing having first andsecond heat transfer chambers operably connected to said first andsecond heat exchanging circuits, respectively, and a rotatable heattransfer structure rotatably supported therewithin about an axis ofrotation, wherein said rotatable heat transfer structure has first andsecond heat transfer portions and a substantially symmetrical moment ofinertia about said axis of rotation and embodies a closed fluid circuitsymmetrically arranged about said axis of rotation and contains apredetermined amount of a heat carrying medium for carrying out athermodynamic-based heat transfer process between said first and secondportions of said rotatable heat transfer structure when said rotatableheat transfer structure is rotated within said stationary housing aboutsaid axis of rotation at an angular velocity within a predeterminedrange of angular velocities; (b) imparting torque to said rotatable heattransfer structure so as to cause said rotatable heat transfer structureto rotate about said axis of rotation and said heat carrying mediumautomatically circulate within said closed fluid circuit; and (c)controlling the angular velocity of said rotatable heat transferstructure within said predetermined range of angular velocities duringstep (b) so that said thermodynamic-based heat transfer process isconducted between said first and second portions of said rotatable heattransfer structure and that heat is transferred between said first andsecond heat transfer chambers.
 58. A method transferring heat betweenfirst and second heat exchanging circuits, comprising the steps: (a)installing between first and second heat exchanging circuits a heattransfer engine which includes a stationary housing having first andsecond heat transfer chambers operably connected to said first andsecond heat exchanging circuits, respectively, and a rotatable heattransfer structure rotatably supported therewithin about an axis ofrotation, wherein said rotatable heat transfer structure has first andsecond heat transfer portions and a substantially symmetrical moment ofinertia about said axis of rotation and embodies a closed fluid circuitsymmetrically arranged about said axis of rotation and having a returnportion which extends along said axis of rotation and has a subportionwith a helical geometry, and said rotatable heat transfer structurefurther contains a predetermined amount of a heat carrying medium forcarrying out a thermodynamic-based heat transfer process between saidfirst and second portions of said rotatable heat transfer structure whensaid rotatable heat transfer structure is rotated within said stationaryhousing about said axis of rotation at an angular velocity within apredetermined range of angular velocities; and (b) imparting torque tosaid rotatable heat transfer structure so as to cause said rotatableheat transfer structure to rotate about said axis of rotation and saidheat carrying medium automatically circulate within said closed fluidcircuit and undergo pressurization as said flow heat carrying mediumflows along the subsection of said return portion having helicalgeometry; and (c) controlling the angular velocity of said rotatableheat transfer structure within said predetermined range of angularvelocities during step (b) so that said thermodynamic-based heattransfer process is conducted between said first and second portions ofsaid rotatable heat transfer structure and that heat is transferredbetween said first and second heat transfer chambers.